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AIR COMPRESSION AND TRANSMISSION 


McGraw-Hill Book Company 


Publisters of Books for 


Electrical World The Engineering and Mining Journal 
Engineering Record Engineering News 


Railway Age Gazette American Machinist 
Signal Engineer American Engineer 
Electric Railway Journal Coal Age 
Metallurgical and Chemical Engineering 


AIR COMPRESSION 


AND 


TRANSMISSION 


BY 


H. J. THORKELSON 


ASSOCIATE PROFESSOR STEAM AND GAS ENGINEERING 
UNIVERSITY OF WISCONSIN 


McGRAW-HILL BOOK COMPANY 
239 WEST 39TH STREET, NEW YORK 
6 BOUVERIE STREET, LONDON, E. C. 
1913 


*,* ; 
- PC 
eater), 


CopyniGnt, 1913, BY THE 


McGraw-Hitt Boox Company 


4 


18 MAY 14 lh 


MoOlurg. 


rchitecture. /F/7A 


PREFACE 


This text is designed to present in logical order the fundamental 


_ principles dealing with the subject of the compression of air, and its 


transmission through ducts or pipes, together with such examples as 
will serve to illustrate their application. 

It is hoped that the presentation will make clear the methods to 
be followed in calculations dealing with air at various pressures, 
and that students and engineers will be better able to appreciate the 
advantages and limitations of the various systems of securing the 
pressures desired, and of using air as a means of transmitting energy, 
or for securing certain results which cannot be obtained under 
normal atmospheric pressure. 

The material offered consists of notes used for a number of years 
by the author in his classes. He wishes to acknowledge his indebt- 
edness to the many excellent texts on the subject which have been 
published, notably those of Richards, Saunders, Hiscox, Harris 
and Peele. 

The fundamental formule are to be found in most texts on Thermo- 
dynamics for Engineers. The author is particularly indebted to 
Prof. C. H. Peabody’s text on this subject, and to the lectures of 
F. W. O’Neill, F. D. Longacre, H. deB. Parsons and F. W. Towl, 
given at Columbia University in a course on Applied Thermo- 
dynamics of Air Compressors. The material on turbo-compressors 
is taken from articles on this subject in recent numbers of the 
Engineering Magazine by Franz zur Nedden. Permission to use 
this material has been courteously granted by these authors and their 
publishers. 

The author is also indebted to the editors of Compressed Air 
Magazine andto various manufacturers for cuts, and to his col- 
leagues, particularly Prof. A. G. Christie and Mr. W. C. Rowse, for 
assistance in preparing this text. 

| Healers 
April, 1913. 


259994 


ae low, A 


RA ee eee ean 
‘ f 7 os 


INTRODUCTION 


HISTORICAL ACCOUNT OF MECHANICAL USES OF AIR 


The earliest writings describing mechanical uses of air are found 
in a book entitled ‘‘Pneumatics”’ by Hero of Alexandria, published 
about 200 B. C. An English translation of this by Bennet Wood- 
croft indicates.a very complete knowledge of many mechanical 
devices possessed by the Ancients, and shows various pumps, 
Hero’s steam turbine and many remarkable uses of air as a means of 
transmitting energy. 


Fic. 1.—Hero’s device for opening temple doors. 


One of the most interesting illustrations is a device for opening 
temple doors by fire on an altar illustrated in Fig. 1. The altar, 
E is hollow and a tube F passes through the altar and is attached 
to a leather bag, K. Beneath this a small weight LZ is suspended 
which is connected to the bag and to the pivots of the temple doors 
as shown. Weights LZ and B are so proportioned as to normally 

Vii 


vil INTRODUCTION 


close the doors. When a fire is lighted on the altar, the bag K 
will expand under the pressure of the heated air below the altar and, 
in doing so lift the weight L. Weight B then falls and causes the 
doors to open. If the fire is extinguished, the air under the altar 
will cool, contract, and the bag K will take the position indicated 
and weight B will be dropped, causing the doors to close. 

A somewhat similar device is also described in the same text 
using heated air under a similar altar to force water from one cham- 
ber into a pail, thus counterbalancing the weight and causing the 
temple doors to open. Many automatons are also described, in 
“which air is made use of to produce musical notes and cause water to 
flow from vessels or objects when certain changes in 
the mechanism take place. 

A-very interesting experiment, illustrated in most 
texts on physics, called ‘‘Hero’s Fountain” and in- 
vented by the author of the book ‘‘ Pneumatics,” 
consists of two globes, Fig 2, M and N, and a brass 
dish, D. The dish D communicates with the lower 
part of the globe N by a tube B and another tube A 
connects the two globes. A third tube passes through 
the dish D to the lower. part of the globe M@. This 
tube having been taken out, the globe M is partially 
filled with water, the tube is then replaced and water 
is poured into the dish. The water flows through 
Fic. 2.—Hero’s tHe tube B into the lower globe and expels the air 

fountain, | Which is forced into the upper globe. The air being 

compressed acts upon the water and makes it jet out 
as shown. If it were not for the resistance of the atmosphere and 
friction, the liquid would rise to a height above the water in the dish 
equal to the difference of the level of the water between the two 
globes. 

Although a knowledge of this wonderful method of transmitting 
energy has been known for centuries, it is only within comparatively 
recent years that it has been used to any considerable extent in 
practical work. Its modern use dates from the construction of the 
Mt. Cenis tunnel completed in 1871. The work on this tunnel, 
which is about 8 miles long, had progressed very slowly from 1857 to 
1861, the tunnel headings having been drilled by hand labor with 
an average advance in each of the two headings of about 1 1/2 ft. 
per day. Machine drills driven by compressed air were introduced 
and the speed rose to 4 3/4 ft. per day, and later when dynamite was 


INTRODUCTION 1X 


introduced, the cut was increased to 6 ft. per day. Sommeiller 
deserves the honor for solving, in this work, many of the initial 
problems of compressed-air production and use. The type of com- 
pressor used is illustrated in Fig. 3. A natural supply of water 
was used for compressing the air. The water was conducted in a 
sluice A through a valve C, compressing the air which is in D and 
forcing it into the reservoir H. When this was done the valve C 
was automatically closed and the water in D allowed to escape and 
be replaced by a new supply of air at atmospheric pressure. This 


Fic. 3.—Sommeiller’s compressor. 


was in turn compressed and forced into the reservoir E, giving not 
a continuous but an intermittent flow of air. This compressor 
furnished air at 80-lb. pressure, but only gave an efficiency of 50 
per cent., that is, only one-half of the available energy was turned 
into useful work. 

Although the value of compressed air for machine drills for tunnel 
work was clearly demonstrated in the building of Mt. Cenis tunnel, 
it was some time before this was applied to mining work. One of 
the earliest tests for mining work was made at the Calumet and 
Hecla copper mine in Michigan in 1878, and the advantages in lower 
costs and higher speeds were so clearly demonstrated that its use 
for this work has since become almost universal. 

One of the most important of modern applications of compressed 
air is to be found in the braking of trains. George Westinghouse, 
in 1869, designed his first “straight air brake,’ which was later 
changed to the “automatic” type of air brake. This apparatus has 


x INTRODUCTION 


been improved and perfected to such an extent that its operation is 
truly marvelous and its application world wide. 

Railroad men were among the first to appreciate the uses of com- 
pressed air in shop and structural work, and its application to 
manufacture and other allied arts has since become so universal 
that a mere recital of the modern applications of compressed air 
would become tedious. One of the largest manufacturers of air 
compressors has recently published a partial list of various ap- 
plications of compressed air for which they have furnished compres- 
sors. This list includes over sixty different industries, with a great 
many different uses of compressed air in each. 

While compressed air has many advantages over other systems 
of transmitting energy, it has also certain disadvantages and limita- 
tions which should be clearly understood. In order to appreciate 
these, it is necessary to study in detail the nature and characteristics 
of air and the fundamental principles governing its generation, 
distribution, and application. 


CONTENTS 


PREFACE. 


INTRODUCTION 


CHAPTER I 


CHARACTERISTICS OF AIR . 


Air—Vapor in air—Free acs aye air—Effect af pressure on aiite 
perature. 


CHAPTER II 


FUNDAMENTAL DEFINITIONS 


Reeth Hine Dewersel cai perakiressA bsolute Renimeracire 
—B.t.u.—Effects of heat—Energy in air—Specific heat—Specific heat 
at constant pressure—Specific heat at constant volume—Real specific 
heat—Apparent specific heat. 


CHAPTER III 


CHARACTERISTIC AND ENERGY EQUATIONS FOR AIR . 


Boyle’s law—Law of Charles—Characteristic equation for perfects gases 
—Numerical value of R—Weight of air—Relation between specific 
heats—Work of isothermal change—Exponential change—Work of 
adiabatic change—Relations between P, v and T for adiabatic and 
exponential change—Computation of intrinsic energy. 


CHAPTER] LV 


GRAPHICAL DIAGRAMS 


AIR 


Construction of isothermal curves—Construction of exponential 
curves—Heat added or taken away for isothermal change—Heat added 
or taken away for exponential change—Difference between isothermal 
and adiabatic compression—Temperatures due to adiabatic com- 
pression—Work done by a compressor—Exponential compression— 
Isothermal compression. 


CHAPTER V 


AT PRESSURES BELOW THE ATMOSPHERE : Ae 
Venturi vacuum pump—Sprengle air Bee ueacurines vacuums— 
Condenser pumps—Wheeler combined pump—Size of water and air 
pumps—Steam cylinder size—U. S. Navy air pumps—Edwards air 
pump—lIndustrial uses of vacuums—Salt evaporating effects—Con- 
centration of liquids—Evaporation of cane juice—Vacuum cleaners— 
Syphon. 
xl 


vii 


Io 


18 


26 


xll 


CONTENTS 


CHAPTER VI 


Arr AT Low PRESSURES 


Uses of air at low Botan eer Me arise sais for ion ea eet ey ce 
forges—Air for cupolas—Air for ventilation—Fans or blowers—Classi- 
fication—Definitions—Measurement of draft—Fan efficiency—Flow of 
gas through an orifice—Loss of head due to friction in ducts—Usual 
velocity in ducts—Notation of symbols—Pipe losses—Rotary blowing 
machines—Blower pressures and capacities—Power for rotary blowers 
—Mechanics of the fan—Effect of outlet on capacity—Work required to 
move a volume of gas—Design°of fans—Description of fans—Centrif- 
ugal fans—Fan blast or steel plate machine—Housing—Cone wheel 
fans—Turbine blast or “‘Sirocco” fan. 


CHAPTER VII 


PISTON COMPRESSORS 


Action of piston Nes ernie tt: eid oF pictow compressor— 
Effect of clearance—Methods of reducing clearance—Suction line— 
Compression line—Wet and dry compression—Actual compression— 
Cards for air compressors. 


CHAPTER VIII 


EFFICIENCIES AND ENERGY COMPENSATION 


Volumetric Efficiency—Apparent volumetric Brrcienmt ernie dave 
metric efficiency—Cylinder efficiency—Efficiency of compression— 
Mechanical efficiency—Net efficiency—Blower efficiency—Energy com- 
pensation—Hydraulic compensator—Lever compensator—Weight com- 
pensator—Straight line compressor—Duplex compressor, 


CHAPTER Ix 


MULTI-STAGE COMPRESSION . 


Advantage of multi-stage Compr eceionen Dress Teed ford various 
stages—Intercoolers—Types of intercoolers—Cooling surface and 
capacity—Intercooler pressure—Effect of clearance on volumetric 
efficiency. 


CHAPTER X 


DETAILS OF PISTON AIR COMPRESSORS 


Classification of valves—Mechanical tee ince valves pete 
Effect of changing discharge pressure—Automatic valves—Valve area— 
Forms of Poppet valves—Piston-inlet valves—Semi-mechanical valves 
—Regulators, unloading devices, etc.—Belt regulator—Westinghouse 
governor—Norwalk regulator—Combined governor and regulator— 
Nordberg governor—Unloading devices—Clearance unloader. 


CHAPTER XI 


TURBO-COMPRESSORS . 


Design of turbo- eerarecenre = alone blower The Rae blower— 


38 


69 


i. 


89 


98 


113 


xX 


CONTENTS 


Cooling turbo-compressors—Cooling devices—Expansion of casing— 
Runners—Balancing axial thrust—Balance by counter-position— 
Balancing by diminishing back area—Balancing by balancing piston— 
Stuffing-boxes—Coupling compressors—Rateau multiplicator—Mixing 
blower. 


CHAPTER XII 


HypDRAULIC COMPRESSION OF AIR Aerie et 
Trompe—Frizell’s renvecccm Baloche end Ncrahniass compressor— 
Arthur compressor—Taylor compressor—-Taylor compressor at Magog, 
Quebec—Taylor compressor at Ainsworth, B. C.—Taylor compressor 
at Victoria Mine, Mich.—Phenomena of hydraulic air compression— 
Losses of hydraulic compression. 


CHAPTER XUHI 


EFFECT OF ALTITUDE AND COMPRESSOR TESTS . 
Effect of altitude on capacity—Effect of mlutuder on arent 
between altitude and volume—Compressor tests—Mode of conducting 
the tests—Results of the tests—Tests of plant No. 1—Test of plant 
No. 2—Test of plant No. 3—Test No. 4—Summary. 


CHAPTER XIV 


RECEIVERS. MEASUREMENT AND TRANSMISSION OF COMPRESSED AIR 
Receivers—Measurement of air and gases—Standards of measure- 
ments—Volumetric meters—Velocity meters—St. John’s meter— 
Venturi meter—Thomas meter—Meter comparisons—Pipe lines— 
Dresser coupler—Hammon coupler—Pipe-line formule—Reheating— 
Stoves. 


Cin b LE Rea, 


THE SELECTION AND CARE OF AIR COMPRESSORS 
Available power—Valve gear—Size and type of eorneeeore Cone 
pressed air explosions—Lubricating compressors—Cleaning valves— 
Inlet connections. 

APPENDIX A—ComMoN LOGARITHMS 

APPENDIX B—-NAPERIAN LOGARITHMS 

APPENDIX C—HyYGROMETRY . 


INDEX 


xlil 


129 


140 


159 


179 


184 
188 
IQI 


201 


254 


>; oF 


So 
74 . 


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a a le vie 
Py St, Sar i 


ee 
Lane 


AIR COMPRESSION AND TRANSMISSION 


CHAPTER I 


CHARACTERISTICS OF AIR 


Air.—Air is a mechanical mixture of several gases, principally 
oxygen and nitrogen, its average composition by volume being as 
follows: 


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By weight it contains about 77 parts of nitrogen to 23 parts of 
oxygen. There may be, in addition to the above, local impurities 
in the atmosphere, the principal ones being ammonia and sul- 
phuretted hydrogen. 

The carbonic acid gas arises principally from the respiration 
of animals and the processes of combustion, but, notwithstanding 
the enormous continual production of this gas, the composition 
of the atmosphere does not vary, for plants in the process of growth 
decompose the carbonic acid, assimilate the carbon, and restore 
to the atmosphere the oxygen, which is being continually consumed 
in the processes of respiration and combustion. 

Vapor in Air.—The vapor of water is always present in the atmos- 
phere and when the air contains as much of this vapor as it possibly 
can, it is said to be saturated. 

The amount of vapor present in the air when saturated will vary 
with the temperature, as shown in Table I, taken from the Smith- 
sonian Institution Reports. 

1 


2 AIR COMPRESSION AND TRANSMISSION 


TABLE I.—GRAINS VAPOR IN 1 CU. FT. OF AIR SATURATED WITH MOISTURE 
(7,000 Grains=1 lb. Avoir.) 


es 4 | I | 2 3 4 5 6 7 8 9 
10 0.481 | 0.505 | 0.529 | 0.554 | 0.582 | 0.610] 0.639] 0.671} 0.704) 0.739 
nme) 0.7767 (02816) 1028504] OF S060) 0.0408 OnOS Sir. 0321.07 Om teal cio meter tou: 
20 Te235e) Le2O4 a) Las S on Le 416 fd 4 Ose eee S 5 0 mn Os ele OO 7imeLeye7 oi eos 
30 T.035' || 2.022 192.1130) 26104 152527900102), 200\e2, 457920550127 040|n2n7A0 
40 2.849 | 2.955 | 3.064 | 3.177 | 3.294 | 3.414) 3.539] 3.667) 3.800) 3.936 
50 4.076 | 4.222 | 4.372 | 4.526 | 4.685 | 4.849] 5.018] 5.191] 5.370] 5.555 
60 52745 | 5. O4L, |) O20A2) 46. C408 On5OS nis On7O27. O00 7 624 Tl 7A COTE 76 
70 7.980 | 8.240\| 8.508 | 8.782 129.066) |.0.356) 02 6551) 0.002) 50.277) 10.007 
80 10.934 |LE275 |Lt 026 111. O87 12.350) (12) 730) tan 27103).520/Ta Os TANenG 
90 I4.790 |15.234 |15.6809 |16.155 |16.634 |17.124\17.626|18.142/18.671|19.212 


This shows the number of grains of vapor present in each cubic 
foot of air when the air is saturated. As there are 7,000 gr. in a 


0 20 40 60 80 100, 
Temperature, Degrees Fahrenheit 


Fic. 4.—Water present in saturated air. 


pound avoirdupois, these weights can be easily converted into 
pounds. This data is shown graphically in Fig 4. 

Free Air.—Free air is air at the pressure and temperature of the 
atmosphere. This is a term used extensively in texts on compressed 
air and in rating the capacity of a compressor. 

It can be shown experimentally (Boyle’s Law) that if a cubic 


CHARACTERISTICS OF AIR 3 


foot of free air at sea-level (14.7 lb.) is compressed to a pressure 
of 44.1 lb. by the gauge, or 4 atmospheres absolute, and allowed 
to cool to the temperature of the atmosphere, the compressed air 
will occupy one-fourth its original volume; if compressed to a pressure 
of 58.8 lb. gauge, or 5 atmospheres absolute, and allowed to cool to 
the temperature of the atmosphere, it will occupy one-fifth of its 
original volume. 

Figure 4 shows the weight of moisture which may be held in 1 
cu. ft. of air at different temperatures, if saturated, and is true 
no matter what pressure the air may be under. 

It is evident then, that if a volume of free saturated air be com- 
pressed into a smaller space and kept at the same temperature, 
part of the vapor it originally contained must be precipitated for 
the reason that 1 cu. ft. of air at a certain temperature can 
only hold a definite weight of vapor when saturated, whether 
compressed or at atmospheric pressure. 

If saturated air is compressed to 5 atmospheres, or 73.5 lb. per 
square inch absolute, and allowed to cool to atmospheric tempera- 
ture, its volume will be reduced to one-fifth of its original volume 
and 1 cu. ft. of compressed air will contain the moisture content 
Obes CU eit, Of iree alr. 


TABLE II—POUNDS OF WATER PRECIPITATED PER CUBIC FOOT OF COM- 
PRESSED AIR AFTER COMPRESSION AND COOLING OF SATURATED 
FREE AIR (Pressures) 


Temp. Ga. 29.4 4AM foSOnO pLO2-07 147.0) 30725 1735 20] 2205.0 
of Abs. 44.1 CO om (aay pO LOL giigole 2: 7A0u 7 | 2210.9 
air Atm. 3 4 5 8 Er 26 5 I51 

fo) .OOOI | .0002| .0003} .0005] .0007| .0017| .0035| .Oo105 
IO .0002 | .0003] .0004| .0008] .ooII| .0027| .0055| .o165 
20 .0004 | .0005] .0007) .0013| .0018) .0045) .0090] .0270 
30 .0006 | .0008) .COII] .0020} .0028) .0070] .0140]} .0420 
40 .0008 | .0012) .0016} .0028} .0040} .0100| .0200] .o600 
50 .0012 | .0017} .0023| .0041| .0058] .0145| .0290] .0870 
60 0016 | .0025| .0033] .0057| .0082] .0205| .o410] .1230 
70 10023 de. CO84) 0045), .0070| JO1I4! .02351 30570" 22770 
80 .003I | .0048| .0062] .o109] .0156] .0390] .0780] .2340 
go SOOA2 she: GOOS4 =. O0GA |) .OT43! (O2TIh 20527. 1055) 23305 


(For a further discussion on moisture in the air see Appendix C.) 


4 AIR COMPRESSION AND TRANSMISSION 


When reduced to the temperature of the atmosphere, the moisture 
held in suspension per cubic foot of the compressed air cannot 
exceed the moisture held in suspension per cubic foot of the free air, 
and in consequence the remaining moisture will be precipitated. 
This will represent for each cubic foot of compressed air a weight 
of water equivalent to four times the weight held in suspension. 

These weights have been calculated for various temperatures 
and pressures as shown in Table IT. 

Dry Air.—Air is said to be “dry” when water evaporates and 
moist objects dry rapidly, and the air is ‘““moist”’ when they do not 
dry rapidly and when the least lowering of temperature brings 
about deposits of moisture. The terms are therefore relative ones, 
but the expression “dry air,’”’? when used with reference to com- 
pressed air, is usually understood as air containing less than half 
the amount of moisture that is contained in “‘saturated”’ air. 

Effect of Pressure on Temperature.—It is well known that air 
can be made to expand by the application of heat. The altar 
trick of the Egyptians illustrates this as does the modern hot-air 
engine. Before friction matches came into general use, fire was 
often produced by means of an air plunger-pump, which consisted 
of a cast-iron barrel weighing several pounds with a bore about 
3/8 in. in diameter, in which a steel piston fitted rather tightly. 
The end of the piston had a small cavity for receiving a piece of 
punk, and by pushing the barrel down on the piston the air in the 
barrel was compressed and its temperature rose high enough to 
ignite the punk. If heat is generated by compressing air, it is 
natural to expect that if compressed air be allowed to expand 
the temperature of the air will fall. This is exactly what does 
happen in compressed-air motors, and if the compressed air contains 
much moisture, the temperature may fall so low that this moisture 
is frozen and collects as a frost in the exhaust pipe. Frost may even 
collect to such an extent as to clog the exhaust pipe and stop the 
motor. ‘The methods used to overcome this obstacle will be dis- 
cussed later in detail. 

The principal characteristics of air to be considered in discussing 
its mechanical uses are: pressure, temperature, volume, weight 
and humidity. 

The relations existing between the temperature and humidity 
have been considered, but before considering the other character- 
istics mentioned, it is necessary to state clearly certain fundamental 
_ definitions. 


CHAPTER II 


FUNDAMENTAL DEFINITIONS 


Work.—Work is a force overcoming resistance, and is measured 
in foot-pounds. A force of 10 lb. exerted for a distance of 4 ft. 
represents 104 or 4o ft.-lb. of work, or a 1oo-lb. weight lifted 
3 ft. represents 300 ft.-lb. of work, etc. 

Energy.—Energy is the ability to do work and may be measured 
in the same units. Energy may exist in a number of forms, for 
example: a water-fall, heat, light, electricity, the wind, etc. The 
source of all energy is the sun, but unfortunately this energy, as it 
reaches the earth, is not in the most suitable form for all of the 
work of the world. It is the province of the engineer to change 
available energy into the desired form with as few losses and as few 
changes as possible. Most of the energy required in commercial 
enterprises is supplied by coal, but the burning of coal represents 
the use of energy from the sun which reached the earth ages ago. 
It is only a question of time when other sources of energy than 
coal will have to be provided in greater abundance than at present 
and the attention of scientific investigators is being called to the 
importance of a more direct method of getting energy from the sun 
and of using other available forms of energy with fewer changes 
than are now necessary. 

Heat.—Heat may be defined as a form of energy, without at 
this time going into any discussion regarding its characteristics or 
effects. 

Power.—Power is defined as the rate of doing work. The engi- 
neers’ unit being the horse-power, or 33,000 ft.-lb. of work per 
minute. 

Temperature.—Temperature is an indication of the direction 
in which heat will flow if it has an opportunity to do so. That 
is, heat will naturally flow from a hot to a cold body. Temperature 
does not represent the heat energy that a substance contains. 

Absolute Temperature.—Temperature in engineering work is 

5 


6 AIR COMPRESSION AND TRANSMISSION 


usually measured on the Fahrenheit thermometer in which the 
freezing-point of water at atmospheric pressure is 32° and the boil- 
ing-point is 212°. As the temperature falls, the vibration of the 
molecules becomes less rapid and the energy contained in any 
substance decreases. That point at which the vibration of the 
molecules ceases is called absolute zero. From a study of the 
property of gases, it is evident that this point is about 460° below 
the zero-point of the Fahrenheit scale. The absolute tempera- 
ture, then, is the sum of the temperature Fahrenheit and 460°. 

B.t.u.—Amounts of heat are measured in British thermal units. 
A B.t.u. is the amount of heat required to raise the temperature of 
t lb. of water from 63° to 64° F. The mechanical equivalent of this 
is about 778 ft.-lb. This is usually represented by the letter J, and 
its reciprocal, or as by A. 

Effects of Heat.—If heat is applied to a substance, many of its 
characteristics may change. Its pressure may change, temperature 
may change; its volume, conductivity, elasticity, etc., may also 
be affected by the application of heat. However, all these effects 
may be classified into two groups: internal changes and external 
changes. This may be represented by an equation as follows: 
Heat applied = internal energy changes + external energy changes. 

The internal changes may be represented in part by changes 
of temperature which mean an increase in the velocity with which 
the molecules vibrate back and forth. This energy expressed in 
heat units may be represented by S. If a substance is of such a 
nature that expansion takes place when heat is applied, then the 
molecules must be separated farther apart against whatever mutual 
attraction exists between them. This also represents an internal 
application of energy and can be represented by the letter L. } 

If external work is done, as in the expansion of any substance 
at constant pressure, this work can be measured by the product of 
the pressure in pounds per square foot and the change in volume 
measured in cubic feet. The product in foot-pounds can be repre- 
sented by the letter W, and its heat equivalent as AW. 

If the heat supplied in B.t.u. is represented by Q, the equation 
Q=S+L+4AW, in which each term is measured in B.t.u. may be 
considered as a fundamental energy equation. 

Energy in Air.—Air may be treated as a perfect gas or as a sub- - 
stance in which there is no mutual attraction existing between the 
molecules. In this case L=O, and the fundamental equation 


FUNDAMENTAL DEFINITIONS © 7 


when applied to air becomes Q=S+AW. That is, if heat is applied 
to air the effect of that heat (Q) will be either to increase its tem- 
perature (S) or to cause the air to expand and do work (AW), or 
both. However, of the heat energy given to the air, the only por- 
tion that can be stored up in the air itself (internal energy) will be 
that portion which is used in increasing the temperature of the air. 
In other words, the internal energy of air depends upon its temperature 
alone. 

This statement is very important and should be thoroughly 
appreciated by the engineer working with compressed air. 

At first thought, it does not seem possible that there is no more 
energy in the air (internal or intrinsic energy) if at atmospheric 
pressure than if the air is compressed and at the same temperature 
as the atmosphere. This, however, is the case, as shown by the 
above equations. 

Although 1 lb. weight of air at a pressure of 1,000 Ib. per square 
inch at the temperature of the atmosphere has no more internal 
energy than 1 lb. of air at atmospheric pressure and temperature, 
still the energy contained in the air under pressure is available for 
use, while that under atmospheric pressure is not, for in the first 
case the compressed air may expand, suffer a loss of pressure and 
also of temperature, cool to a point below the temperature of the 
atmosphere, and in that way give up a portion of its internal energy. 
The greater the fall of pressure during expansion, the greater the 
fall of temperature, and hence the greater the amount of internal 
energy available for use. 

Some engineers are under the impression that the energy used in 
compressing air is actually stored up in the air. This, however, is 
far from true, the internal energy in compressed air depends on its 
temperature alone, and that part of this internal energy that may 
be available for use will depend upon the fall of pressure and hence 
of temperature that is permissible. 

Specific Heat.—The specific heat of a substance in English units 
is the amount of-heat required to increase the temperature of 1 |b. 
of the substance by one degree, and is usually represented by C. 

Specific Heat at Constant Pressure.—The specific heat at con- 
stant pressure (Cp) is the amount of heat required to increase the 
temperature of 1 lb. of the substance one degree F. the pressure 
remaining constant. 

Specific Heat at Constant Volume.—The specific heat at constant 
volume Cy is the amount of heat required to increase the tempera- 


8 AIR COMPRESSION AND TRANSMISSION 


ture of 1 lb. of the substance one degree, the volume remaining 
constant. . 

As external work is done during a change at constant pressure, 
it is quite clear that the former specific heat is greater than the 
latterthatusy-C p. eC: 

Real Specific Heat.—The real specific heat of a substance is the 
amount of heat required to merely increase the temperature of 1 lb. 
of the substance, one degree F. that is, this excludes any energy 
that may be used in doing external or other work. 

Apparent Specific Heat.—The apparent specific heat is the amount 
of heat required to increase the temperature of 1 lb. of the substance 
one degree F. including heat used in doing external or other work at 
the same time. This is, therefore, usually greater than the real 
specific heat. From the fundamental equation Q=S-++AW, it is ap- 
parent that, if all the heat applied is to be used in raising tempera- 
ture, AW =O and Q=S. This condition can only exist if there is no 
change in volume. 

For a perfect gas, and hence for air, the real specific heat is 
equal to the specific heat at constant volume, that is, Co. 

These specific heats are measured in heat units, but may be ex- 
pressed in foot-pounds by multiplying by the mechanical equivalent 
of a heat unit or 778 ft.-lb., or J. When this is done, the specific 
heat is represented by K, that is, 


AED and JC p=K »p. 


TABLE III.—C, FOR AIR AT VARIOUS PRESSURES AND TEMPERATURES 


Pressures in atmospheres and pounds per square inch 
absolute 
Temperatures 
Fahrenheit 

I 10 20 40 70 100 
LAND eet A TLDs 294 lb. 588 lb. | 1,029 lb. | 1,470 lb. 
212° On2202 0. 2389 0. 2408 0.2446 Oresr2 0.2583 
Bo ONz275 O.2419 0.2465 O12 512 0.2773 0. 2986 
—58° ON238O0umO. 2455 3. 0.25720 (One 755m mons TOMImOnd laa 
— 148° 07238010) FO. 2585" 1. °O,, 2004 mt O es 00 MMO eA OT ie ce meee 
— 238° O72424 Cl OAS TOS Ul GORSO4S Arr ee aia ie eee tent 
—274° 02240791 AOL 4147 Spree. ie meee aeaegp noid eee ae sire ee 


FUNDAMENTAL DEFINITIONS 9 


The specific heat of air at constant pressure, (C p) is usually taken 
as 0.2375 B.t.u., and the specific heat at constant volume, (Cy) as 


; : G 
0.1689 B.t.u. The ratio of these two specific heats, Ce. Or 1.405; 
Vv 


is frequently used as shown later in compressed-air calculations. 

As a matter of fact, the specific heat of air at constant pressure 
is not the same under all conditions, as it increases with increasing 
pressures or decreasing temperatures, as shown in Table III given by 
Prof. Linde. 


CHAPTER III 


CHARACTERISTIC AND ENERGY EQUATIONS FOR AIR 


The principal characteristics of air to be studied are its pressure, 
volume and temperature. Pressure may be measured in pounds 
per square inch or pounds per square foot, and in the equations that 
follow, absolute pressure is used. This is the sum of the pressure 
of the atmosphere and the pressure shown by the gage. When 
measured in pounds per square inch, the pressure will be indicated 
by p. When measured in pounds per square foot, it will be indicated 
by P. Volumes are measured in cubic inches or cubic feet, usually 
thelatter. If the volume of a1 lb. weight in cubic feet is considered, 
it will be represented by v. If the total volume of any weight in 
cubic feet is considered, it will be represented by V. Temperature 
may be measured on the Fahrenheit scale, and when so done is indi- 
cated by ¢. If absolute temperature is considered, it will be repre- 
sented by 7. TF-=t-+-460. 

Boyle’s Law.— Boyle’s law for perfect gases, which was determined 
by experiment, states that the product of the pressure and volume of 
a perfect gas is a constant if the temperature is constant. ‘That is, 
PV1= pod2= p3ds, etc., if the temperature is constant, or PiVi= 
PoVeo, etc, Or P3,=P oto. 

Law of Charles.—The law of Charles states that the volume varies 


inversely as the temperature, if the pressure is kept constant. Or 
Vie Ce : ; 
m =7, etc., if the pressure is constant. 
Leads 

Characteristic Equation for Perfect Gases.—Combining these two 
Piri uN Povo 
be T > : 


a constant is obtained. This constant will be indicated 


equations, which were determined by experiment, or 


Bal ie 
Lael oe 
as R and the above equation as Pyv,:=RT}. 


Numerical Value of R.—The value of this constant (R ao 7) will 
1 


vary with the gas considered. For example: As 1 lb. weight of oxygen 
at 32° F. and 14.7 lb. absolute pressure occupies 12.21 cu. ft., the 
| 10 


- 


CHARACTERISTIC EQUATIONS FOR AIR 1a 


14.7 X144 X12.21 
reais eA 
In the same way, as the volume of 1 lb. weight of air at this tempera- 
ture and pressure is 12.39 cu. ft., the value of this constant for air is 
14.7 X144 X12.39 | 
460+ 32 mice 
1V1 _ Pode 


Weight of Air.—In the equation, “Pin i meeepierne dey Keven a, 
this may be multiplied by the number of pounds of air, giving the 
equation SA aA E ss, 3, in which w represents the number of 
pounds of air. This equation is of great assistance in determining 
the weight of air in a certain receiver. For example: If a receiver 
3 {t. in diameter by ro ft. high, and having, therefore, a volume of 
70.68 cu. ft., contains air at a temperature of 70° F. and a gauge pres- 
sure of 143.3 lb., or 158 lb. absolute, the weight of air contained in 
the réceiver if the pressure of the atmosphere is 14.7 lb. per square 
Pr vie tso 144 70.08 
53.321 §3-3X(460+70) 9°91” 

Relation between Specific Heats.—If 1 lb. of air contained in a 
vertical cylinder having a weighted piston above is heated, the 
temperature of the air will increase and, as the pressure is kept con- 
stant, the heat absorbed, may be expressed as the product of the 
amount of heat required to cause a change of one deg. F. and the 
actual change in temperature measured in degrees F. or as 
CONC intB tiesony ke (haa) eit expressed in mechanical 
units; where 7; and J, are the initial and final temperatures 
respectively. 

S, the energy required to cause the increase in temperature alone, 
is Cy(T2—T}) in heat units; K,(T2—71), if expressed in foot-pounds. 

W, the external work, must be P:(v2—21), where Pj represents the 
pressure on the piston in pounds per square foot; v1 and v2 the initial 
and final volumes of the air respectively in cubic feet, but as Pu= 
RT this may be expressed as R(T2—T7)); that is, the heat energy ap- 
plied, or 77830=K.(T2—T1)+R(T2—T;), in foot-pounds. But the 
heat energy applied may also be represented by the expression, 
K »(T2—T71), hence: 

K »(T2—T3) =K,(T,—-T,)+R(T2—-T1) 
K p=K.+R, or R=K p—Ky=778(Cp—Cv) = 
778(0.2375 —0.1689) = 53.3 
If expansion of air takes place in a perfectly non-conducting 


value of this constant for oxygen will be 


inch will be 


12 AIR COMPRESSION AND TRANSMISSION 


cylinder, Q of the equation Q=S+AW must equal O, and hence 
W= — or W = —7785; that is, during an adiabatic expansion of air, 


as it is called, the temperature falls and the amount of energy used 
in doing work during the expansion will be given by the expression 
K,(T1—T>2) in foot-pounds. 

As no heat energy is given to the substance nor taken from it 
during this change, all the work that is done must be done at the 
expense of the internal energy. 

The principal relations known to exist between Kp and Kv for 
air are as follows: 


K» he, eee 


K, 
K p=KytR=Kot53-3 
K p—Kv=53-3 


Work of Isothermal Change.—From the characteristic equation 
for air, it is evident that if the temperature is a constant, PV = 


?P 1 


PV= RPV, = PVo 


Fic. 5.—Isothermal change of air. 


constant, or PiVi=P2V2=P3V3, etc. That is, if a certain volume 
of air is compressed to one-half its original volume isothermally, 
its pressure will be doubled, while if its volume is reduced to one- 
fourth its original volume, its pressure will be quadrupled, etc. 
This relation follows the path shown by the isothermal curve of 
the chart, Fig. 5, the equation being that of an equilateral hyperbola. 

In order to find the work done during an isothermal compression, 
it is necessary to find the area under the compression curve drawn | 
on a PV diagram or plane, this area representing the work done 


CHARACTERISTIC EQUATIONS FOR AIR 13 


expressed in foot-pounds. ‘The fundamental expression for any area 
on a pressure—volume diagram is given by the expression, 


Area=work = { PdV 


with the isothermal curve PV =P1V,, hence pan, and 
V2 Ve 
w= { Pav ea | dV 
V V 
1 Vi 
Ve. 
hence, W =P1V, loge = 
Vi 


where P, represents the maximum pressure in pounds per square 
foot and V; the minimum total volume of the compressed air in 
cubic feet, V2 the total volume in cubic feet occupied by the air 
before compression. 

This also represents the work done by any number of pounds 
of air expanding at constant temperature from an initial volume 
of V; cu. ft. to a final volume of V2 cu. ft. If 1 Ib. of air expands 
in this way the work done will be represented by the equation 


P21 loge = or as P\vi= RT), this may be written RT; loge . 
As the relation P1v,;=P2v2 applies to this isothermal change, 
| Py 
2 Ps 
To illustrate the application of this formula, suppose it is required 
to find the work done as 1 lb. of air expands at constant temperature 


of 120° F. from a pressure of 150 lb. per square inch absolute to 
30 lb. per square inch absolute the work done will be: 


: ae 
the equation may also be written Pv; loge si or RT, loge 


ISOX144 
53.3X(120+460) loge peat 


It is evident that the ratios of pressures in pounds per square inch 
is the same as the ratios of pressures in pounds per square foot. 


53-3 X 580 Xloge 5 =53-3 X 580 X 1.6094, or 49,800 ft.-lb. 


A table of logarithms to the base e, or hyperbolic or Naperian 
logarithms as they are called, is given in Appendix B. This is 
for numbers from 1 toro. If it is desired to obtain other logarithms 
as for the number 0.12, this is equal to the loge of 1.2 minus the loge 


14 AIR COMPRESSION AND TRANSMISSION 


of ro, or the loge of ee In the same way the loge of 25 is the same 


as the loge of (2.5 X10), or the loge of ro plus the loge of 2.5. 

Work of Exponential Change.—When the equation of the com- 
pression line is unknown, it may be represented by the equation 
PV”=a constant, where z is the unknown exponent (Fig. 6). 


F 1 


Pyns PV, "= PV" 


Fic. 6.—Exponential change of air. 


PV"=P,V," and par Substituting in the formula for 
work 
Ve V2 V2 
: d 
w= frav,we | Davie gen | poner | V-"dV 
Vi 1 1, 
n artes Ln 
pas ee Mee as Pi Va Pa Vere 
P2Vo—PiVi_ PiVi-—PoVo2 


I—n N—TI 

P, represents the maximum pressure and P, the minimum pressure 
in pounds per square foot, and V; the minimum total volume of 
the compressed air in cubic feet and V2 its volume in cubic feet 
before compression. 

The above expression for work also represents the number of 
foot-pounds of work that will be done by any number of pounds of 
air expanding according to a change of pressure and volume repre- 
sented by Pai =LfsV 5". 

If 1 lb. of air expands in this way the work done will be represented 
by the equation 
ya P ate _ ROT 2). 

n—I N—I 


As Pyvy" =P vq", Pove=P 01 (°:) 


V2 


CHARACTERISTIC EQUATIONS FOR AIR 15 


From this it is evident that the equation may also be written: 


Pv (*) ‘| Piven es) =| 
SLY mes he > or as Tle 
nN—-TI V2 n—I ) 1 
As an illustration suppose it is required to find the work done 
as 1 lb. of air expands according to the equation Pyv\)*=Po2""4 
from an initial pressure of 180 lb. per square inch absolute and a 


- volume of 1.2 cu. ft. to a final pressure of 15 lb. per square inch 
absolute, the work done will be 


oe ga | | 
ee eres = I—0.083 


ye Neem 180 AT 
=77,760[1 —0.492] =77,760 X0.508 = 39,502 ft.-lb. 


Work of Adiabatic Change.—If a change takes place in a cylinder 
surrounded by non-conducting walls preventing energy in the form 
of heat from entering or leaving the cylinder, this change is called 
adiabatic. , 

During an adiabatic expansion of 1 lb. of air, the work done may 
be represented by the equation Ky(71—T+2), as for such a change 
AW =—S or W=—778S, and S=C»(T2—T)). 

As Pa Rt and Po,=RT > 
K (P01 —P 202) 


K»(%i—T,)= R 


but R=K p—Ky 


KyPii—Pod2) _ P1i1— P22 
GOK Kp 


Kee 


This expression must equal the expression for work under an ex- 
ponential curve, viz.: 


then Ky (T,—T2) = 


Pi1—Pove 
n—I 
hence for an adiabatic change of 1 lb. of air, the exponent in the 


equation P11” =P v2” must equal —; that is, 


Ky 
Kp Kp 
Py, Ke =PypKo and for w lb. 
Kp Kp 


PV Ko PV Kk» =PsV; acae 


16 AIR COMPRESSION AND TRANSMISSION 


Relations Between P, v and T for Adiabatic and Exponential 
Change.—This shows the relations existing between P and v for 
an adiabatic change. In order to find the relations between Pand 
T or v and T for such a change, it is necessary to turn to the charac- 
teristic equation Pu=RT, or 


P44 Pods P3033 


Ty ah T 2 cs T3 
Ey peels 
then PRE 
But as P4013-49 = Povo! 495 
i ‘- 2) 1+405 
then es & 
: VN om Oal (**) 0-405 Ty 
Equating ("*) Sais and : Te 
vas © 1.405 
As Fat 
Pi) van ey ie nao gph 
ened Bee teeth Vo . h ) 1.405 2 
then & = ence : T, 
(a 28825 Tae (*) 0.405 
Or, P, te - 


In the same way for changes represented by the exponential 
equation P41” =P2v”. 


("2 pieaae eb S eee 
V1 CoN De ae P» RS 


Computation of Intrinsic Energy.—These equations enable 
calculations to be made of temperature changes during adiabatic 
compression and are used in calculating the heat curves of the chart, 
Fig. ri. 

During an adiabatic expansion of air all the work that is done 
must be doneat the expense of the intrinsicenergy. This can be shown 
from the equation Q=S+AW, which becomes O=S+AW for an 
adiabatic change, for from the definition Q must equal zero. This 


being true, Weare or a measure of the work done would give also 


CHARACTERISTIC EQUATIONS FOR AIR 17 


a measure of the change in intrinsic energy. That is, if the initial 
pressure and volume of w Ib. of air is known, the equation 
Eee overeat) too VOW tet 1h oP 


ToL IeA0 Saat 0.405 


will be a measure of the intrinsic energy available. In this equation 
P, and P2 represent respectively the initial and final pressures in 
pounds per quare foot. 

If v; and ve be used representing the specific volumes, that is, 
the volumes occupied by 1 lb. weight, then the equation given 
represents the work done and the change of internal energy of 1 lb. 
weight only; but if V;and V2 be used representing the initial and final 


volumes occupied by all the air concerned in the expansion, the 
PiVi—P2V2 Fhe ° ° : ° 

gee _will represent the entire amount of intrinsic 
energy available for use in expansion of w lb. from:pressure P; and 
volume V; to a pressure P, and volume V2. 

If P; and V; are known as well as Po, the final pressure, V2 may be 
calculated if the equation of the expansion line is known. ‘This 
equation will be 


equation 


P1V 434% = PoVo! 4 


the equation for an adiabatic expansion. 

P\Vi-—PoV2 
0.405 
cannot be used for calculating a change of internal energy. In 
order to obtain the amount of internal energy contained in air it is 
merely necessary to assume an adiabatic expansion to infinity. 


The area under such an expansion curve is finite and amounts 
PiV, 


If the expansion is not adiabatic, the equation 


to ate For example, the internal energy contained in 1 ID. of air 
405 
at atmospheric pressure and 32° F., with a volume of 12.39 cu. ft. is 
P40; RT, ; 
eee. OF , which is 
0.405 0.405 
TBE r 53:2%493 or 64,730 ft.-lb. 
0.405 0.405 


It is, of course, impossible to obtain this amount of energy from a 
pound of air, as it is impossible to secure its complete expansion to 
absolute zero of pressure. 


CHAPTER IV 


GRAPHICAL DIAGRAMS 


Construction of Isothermal Curves.—It is frequently necessary 
or desirable to construct graphically compression curves represent- 
ing isothermal and adiabatic changes of air. A method of con- 
structing the isothermal curve is shown in Fig. 7, in which O rep- 


ie a 1 4 b 


Vv 


Fic. 7.—Graphical construction of equilateral hyperbola. 


resents the intersection of the coordinates of a pressure-volume 
plane. If an isothermal line is to be drawn through point a, con- 
struct horizontal line a—b and vertical line a—c, as shown, then 
draw any diagonal line as o-1 and complete the rectangle a—1—2-3; 
3 isa point on the required curve. In thesame way, other diagonals, 
as o-4, may be drawn, and the rectangle a—4-5—6 constructed, 
giving point 6 as another point on the required curve, etc. 

Another method of constructing this curve is shown in Fig. 8. 
Assume that it is required to draw the isothermal] line for air through 
point a. A diagonal line, as 1-2, may be drawn through this point 
and the distance 2-b made equal to 1~-a; 0 is a point on the required 
curve. In the same way line 3-4 may be drawn through 0 and the 
distance 4—c made equal to the distance 3-0, giving another point c 
on the required curve, etc. 

In the equation PV”=P,V1"=Po2V.", etc., the value of the 
exponent may be found, if the values of P and V for any two points 

18 


GRAPHICAL DIAGRAMS 19 


are known, by taking logarithms of both sides of the equation, 
PiVi"=P2V 2" 

log PitnXlog Vi=log Pe+n Xlog Ve, and from this 
_ log Pi—log Pe 


"log Vo—log V1 


Fic. 9.—Graphical construction of exponential curve. 


Figure 9 represents graphically a curve whose equation is P}V,"= 
P2V,” in which the numerical value of 7 is 1.405. Curves of this 
type may be classed as exponential or logarithmic curves. A 
simple method (Brauer’s) of constructing such curves graphically 
is given below, together with a development of the equations used. 

Construction of Exponential Curve.—Brauer’s method of con- 
structing an exponential curve may be illustrated by assuming any 
two points, as A and B, Fig. 9, of such a curve and drawing lines 
through both points perpendicular to both axes. Through C and £ 


20 AIR COMPRESSION AND TRANSMISSION 


draw lines making an angle of 45 degrees the withaxes. D represents 
the intersection of CD with BH produced, and F the intersection of 
EF with AL produced. 

Connecting points D and F with O will give the two angles DOP 
or 8 and FOV or a. In order to determine the relations between 
these angles, the following demonstration is given: 


PPG = Pe GH = eee 

DH =P, tan B 

P, =P ,(1+tan P) (1) 

Vez =s V,ztGB= V,tLE= V,+tiPF 

ine = Ve4 tan @ 

Ve =Va(ittan a) 

Ve=V4(rt-+tan a)” (2) 

Multiplying equations (1) and (2) 

PzV%(1+tan a)" =P,V3(1+tan £) 
but, Puvi Pave 

(rt+tan a)"=1-+tan 6 

tan S=(1-+tan a)"—1 


This shows the relation between the angles § and a in terms of the 
exponent 7. 

For convenience, it is customary to make tangent of angle a= 
0.25. Tan ? has been computed for various values of 7 as follows: 


n Tang. 2 
O.7 0.169 ® 
0.8 0.195 
0.9 C5223 
iO On25 
1.0646 0.268 
TESS 0. 288 
r25 Ou322 
I. 333 0.347 
T3239 0.358 
TAt One7 


If, for example, a curve is to be drawn through any point, as A 
following the equation PiVi17>=P2V_!>, lay off angle VOF, or a, 
with a tangent equal to 0.25 and angle POD, or , with a tangent 
of 0.322; then draw AC and AL through A, and construct Jines CD 
and FH, making angles CDH and FEL equal to 45 degrees, the in- 
tersection of the horizontal line from D and the vertical line from £ 
will give point B as one of the required points of the curve. In the 
same way, other points of the required curve may be obtained. 


GRAPHICAL DIAGRAMS 21 


Heat added or taken away for Isothermal Change.—The funda- 
mental equation showing the effects of applying heat to air, Q0=S-+ 
AW, enables a determination of Q, the heat to be applied in the case 
of expansion, or the heat to be taken away in the case of compression 
in order to cause the expansion or compression line to follow a certain 
exponential curve. If this required expansion or compression curve 
is to be isothermal, S=O and Q=AW, that is, the heat to be applied 
to secure isothermal expansion must equal the heat equivalent of the 

2 


. P1Vi loge Vi . 
work done during the expansion, or — cae * heat units, and in 
the same way the heat to be taken away, in order to secure isother- 
mal compression, must equal the heat equivalent of the work done 
during the compression. 

Heat Added or taken away for Exponential Change.— Frequently 


it is desired to secure a compression or expansion curve between the 


P 


0 


Fic. 10.—Graphical measurement of change of heat energy. 


isothermal and adiabatic, following the equation PaVa"=PsVo", as 
shown in Fig. 10, in which z is less than 1.405. In this case Q= 
S+AW. It has been pointed out that the mechanical equivalent 
of the internal energy possessed by air is represented by the area 
under an adiabatic curve continued to infinity and in Fig. 10 af 
represents such a curve through point a and be such a curve through 
point 6. W, or the external work done between a and 8, is repre- 
sented by the area abcd, or 1+ 2. SX778, or the internal energy for 
point a by the area 1+4, and SX778 for point b by the area 3+-4. 
The change of internal energy in going from a to 6 in the case of 
expansion will be area (3+4) —area (1+4), or 3—1, and as W= 
1+2, the mechanical equivalent of Q must equal SX778+W, or 


22 AIR COMPRESSION AND TRANSMISSION 


(3-1) +(1+2), or (3-+2). That is, the mechanical equivalent of 
the: heat energy to be added during expansion from a to b, or 
taken away during compression from 6 to a, in order to follow a 
certain curve on the PV diagram, will be the area bounded by the 
curve ab and two adiabatics from the ends of the curve to infinity, 
PuVa—PoV5 Pov P pV, 
n—I By 0.405 - 0.405 

Difference between Isothermal and Adiabatic Compression.— 

The difference between adiabatic and isothe:mal compression is 


or, 


367.5 


3528 
333.1 
3234 


EEE 


CATR 


aang 
ae NSE SN NIN 


ved a 


Sl Orme x 
af COCCI el Ks 
gms a EOLA NRE 
BREE SCOUDSENEN: 
ee gel LECCE 
eo CELLET ALIS NINE 
eee TILLELLLELLLL ALAN NINE 
gree HCL TAIN De NINE 
ep ALESIS CURSES: 
breeg CILCCE EEC Net NN e 
serge HCL CTT TTS SEINE 
deo F ICCC SSE SSE 
3 1323 10 Ps S ; °D 
em eC LCLLEL LLC NEN 
o 102.9 o vie PS PRR NS EPS ‘a 
2 ore COSC Ge seel: 
pes Cee TTS ie ESSE 
58.8 ears Ss 5 
441 4 Hee eee 3 
29.4 CREE BEE EE i 
)4e/ ec Stipes Senses 
e Haan 0 76 Seah 
02 ean niet fen eae eres 


Fic. 11.—Temperature change due to adiabatic compression. 


illustrated by the two curves on the PV plane at the left of Fig. 11, 
which shows that while an isothermal compression of free air to half 
its original volume will raise its pressure to 2 atmospheres, or 14.7 


GRAPHICAL DIAGRAMS 23 


lb. gauge at sea-level, the same reduction of volume adiabatically will 
raise its pressure to 2.82 atmospheres, or 26.75 lb. gage. Jf the re- 
duction of volume is to 0.2 of the original volume, the pressure for 
isothermal compression will be 5 atmospheres or 58.8 lb. gage at sea- 
level, and for adiabatic compression 8.88 atmospheres or 115.83 Ib. 
gage. 

Temperatures due to Adiabatic Compression.—Adiabatic com- 
pression is always accompanied by an increase of temperature fol- 

0-405 0-2882 
a FZ) . ots Gs). This shows the 
ratio of the final absolute temperature to the initia! temperature. 

The right-hand part of Fig. 11 shows the resulting temperature 
Fahrenheit for adiabatic compression with initial temperatures vary- 
ing from 0° to 100° F._ From this it is evident that adiabatic com- 
pression to 0.2, the original volume and consequently 8.8 atmos- 
pheres, will approximate a final temperature 510° F. if the initial 
temperature is 60° F. 

Very high temperatures are to be avoided in compressing air as an 
explosion may result if the temperature is sufficient to ignite the 
volatile matter contained in the lubricating oils used. In addition 
to this, the energy represented by the high temperature will soon be 
dissipated by radiation. Isothermal compression requires removal 
of heat energy during compression, and this cannot be accomplished 
satisfactorily with piston or fan compressors operating at modern 
speeds. Because of these conditions, high compression is secured in 
modern compressors by compressing by stages and cooling the air 
between stages. Compression in single-stage compressols Is usually 
accompanied by such cooling as can be secured by water-jackets or 
other means, but with the speeds required the usual effect is to se- 
cure a compression curve with an exponent m varying between 1.25 
ance. 32" | 

Work done by a Compressor.—The work done in a machine, which 
draws in air, compresses it, and then discharges the compressed air, 
can be calculated, it the suction and discharge pressures are known 
and the character or exponent of the compression curve is given. 

Exponential Compression.—Let Fig. 12 represent such a series of 
changes for any compressor, in which the effect of clearance is dis- 
regarded. In this diagram, which is somewhat similar to the indi- 
cator card from a piston compressing cylinder, d—a represents the 
intake of air at a pressure of p2 Ib. per square inch, a—d represents 
the compression of this air from p2 to pi Ib. per square inch following 


lowing the equation 


24 AIR COMPRESSION AND TRANSMISSION 


the compression curve p1V 1" = p2V 2” anb-c represents the discharged 
of the compressed air at a pressure /; lb. per square inch. 

The area enclosed by the lines d-a—b-c-d will represent the work 
done if V represents the volume in cubic feet and if p, representing 
the pressures in pounds per square inch, be multiplied by 144 to give 
pressures in pounds per square foot. 


Pp 


—— eo Pimper oe 


Fic. 12.—Work diagram of air compressor. 


The required area will be area a-b-g-f plus area b-c-e-g minus 
area a—d-—e-f, 
boVo—paVa 
nan alae 


or 144(poVo—paVa) (+1) ) or —-144(oV>—PaVa) 


Va is usually known in compressed-air calculations, but Vz is not 
known directly, but as paVa"=poVo” this equation may be sim- 


or +144 poVo—144 PiVa 


n—-1 es 
plified, for p»Vo=paVa (7) > or poVo= paVa &) m , andethe 


expression may be written: 


EDN (ere it | | 
tees I 
Suppose, for example, it is desired to ascertain the work required 
to compress 2,500 cu. ft. of air from a suction of atmospheric pressure 
or 15 lb. absolute to a gauge pressure of 100 Ib. per square inch, or 
115 lb. absolute, following a compression Jine whose exponent is 1.3. 
1-3-1 


rhe) ‘ I.3 (=) j.2o0e 
This will require 7S xX 244X 152,500 | me : 


0-23 
oy 4.33 X144X15 X 2,500 (7.6) — 4 


or, 23,375,000 X (1.59—1), Or 13,780,000 ft.-lb. 


GRAPHICAL DIAGRAMS 25 


If the compressor is to have this free-air capacity of 2,500 cu. ft. 
per minute, the horse-power required will be 
13780000 
33000 


SOL ATO. p), 


Isothermal Compression.—If the compression line a—b was an 
isothermal line, following the equation paVa=poV», the expression 
for the area would be 


ie Vig 
144 ov loge V+ Ve baVal, or 144 poVo loge +, 


or, 144 paVa loge ae Or 144 X15 X2,500 loge a 


a 
Or, 5,400,000 loge 7.66 = 5,400,000 X 2.036 = 11,000,000 ft.-lb. 
If the compressor is to have this free-air capacity of 2,500 cu. ft. 
per minute, the horse-power required will be 


I I000000 
33000 


Or, 233 hep: 


These calculations show the advantage of having the exponent 
of the compression line as low as possible, or in other words keeping 
the temperature of the air during compression from rising. The 
advantages and methods of attempting this are discussed later. 

The effect of valves, clearance and friction on the required horse- 
power is considered later in discussing piston compressors. 


CHAPTER V 


AIR AT PRESSURES BELOW THE ATMOSPHERE 


A study of the properties of air, and of its applications would 
not be complete without reference to at least a few of the uses of 
air at pressures below the atmosphere. 

For purposes of experiment and for laboratory uses, these low 
pressures are usually obtained by means of the familiar air pump. 

Venturi Vacuum Pump.—Another method of securing these low 
pressures is by means of a very simple hydraulic air ejector or 
‘“‘venturl vacuum pump”’ as it is sometimes called. 

This convenient instrument for quickly obtaining an approximate 
vacuum depends on the principle that a fluid passing at a high veloc- 
ity through a converging and diverging nozzle in which the curves 

conform to the shape of the “vena 
contracta”’ of a jet from an orifice, 
will produce an approximate vacuum 
at a point nearest its greatest con- 
traction and if an air chamber is 
connected through an orifice at 
this point the air will be drawn 
into the jet and a very good 
vacuum formed in the chamber. 
In the sketch shown in Fig. 13, 
tube A may be connected by a rubber hose to a faucet. The con- 
verging-diverging tube through which the water is forced is shown 
at D. Tube C, which is connected with the chamber from which 
the air is to be exhausted, has a check valve E and is connected 
to the smallest diameter of the nozzle. It has been found that 
better results are secured when a baffle F is introduced into the dis- 
charge pipe B, as shown. 

Sprengle Air Pump.—For a more perfect vacuum than the air 
pump or the hydraulic air ejector the Sprengle mercurial air pump 
is used. This pump depends on the fact that if mercury is forced 
through an inverted U tube the mercury going over the bend will 

26 


Fic. 13.—Hydraulic air pumps. 


AIR AT PRESSURES BELOW THE ATMOSPHERE 27 


exhaust the air from any chamber that is connected to the U tube 
at the top of the bend. 

This pump is used for exhausting the air from incandescent 
lamp globes and remarkably low pressures are secured with it. 

Measuring Vacuums.—Although the normal pressure of the 
atmosphere at sea-level is 14.7 lb. per square inch and pressures 
above that are designated in the same units, pressures below the 
atmosphere are not usually so designated but instead are expressed 
in inches of mercury. If a U tube at the sea-level is filled with 
mercury and one end connected with a perfect vacuum while the 
other is in contact with the atmosphere, the mercury will rise to a 
height of 29.92 in. above the level of the mercury in the leg that is 
exposed to the atmosphere (1 in. of mercury=o.49 lb. per square 
inch). Consequently a vacuum gage indicating 20 would represent 
two-thirds of a perfect vacuum or a pressure of about 10 lb. below 
the atmosphere, that is, approximately 5 lb. absolute. 

One of the most familiar uses of pressures below the atmosphere 
is in a condenser for steam engines in which the back pressure of the 
engine is reduced considerably below that of the atmosphere, thereby 
increasing the power of the engine. 

Condenser Pumps.—Condenser pumps! are of two kinds: cir- 
culating pumps and air pumps, the circulating pumps being used 
to force the condensing water through the condenser and the air 
pumps for removing the condensed steam and air. In some types 
of condensers, the condensed steam is removed by gravity, as in the 
barometric type, and the air pump removes but air alone, being 
in this case called a ‘‘dry-air”’ pump to distinguish from the “‘wet- 
air’? pump, which removes condensed steam as well as air. 

Wheeler Combined Pump.—As a circulating pump _ usually 
lifts the water but a short distance, it is built as a tank pump, but 
should it be required to lift water through a long line of pipe, as a 
line to the top of a cooling tower, it must be made of heavier con- 
struction. Because of the large quantities of water which it handles, 
it is very frequently built of the centrifugal type. Fig. 14 shows a 
combined air and circulating pump of the Wheeler Condenser and 
Engineering Company. The steam cylinder is in the center, the 
air-pump cylinder at the left and the circulating-pump cylinder at 
the right. | 

The circulating pump forms one end support for the condenser. 
The water is discharged through A into one set of tubes and then 


1 Pumping Machinery, Greene. 


28 AIR COMPRESSION AND TRANSMISSION 


it returns through B and the upper set of tubes to C, where it dis- 
charges. The air pump forms the other support for the shell. 
It takes the air and water from the condenser and discharges it 


through D. The suction space F is connected to G. 


et 
SSSSSPSSSSUAY 4 
Vee pies 


PN 


S SASS SSS 


Sixes 
an 


YN 


N 
G, 


rss) SS 
jp PIE GELS z 


Ve 
EA suk NOSE 


seal 


ISspsmsssssJ 


=} NS 
GLLLLEL peer 
Zz 


Air and Condensed Steam Stearn Cylinder Circulating Pump 


is) 
LELLLN 


NN CiLLL/ LLL, 
WA 
TSS N 


Fic. 14.—Wheeler condenser pump. 


Size of Water and Air Pumps.—To find the size of the water and 
air ends of the pump, suppose that W pounds of steam per hour at 
a pressure p are to be condensed. If 7 is heat of vaporization of 
the steam, x its quality, ¢.° the temperature of the condensed steam, 
and g the heat of the liquid, and if G pounds of water entering at 
ti° F. and leaving at ¢.° F. are to be used, G is given by the equation: 


5 W (q+%.r— ic) 
dto—Qti 
in which the subscripts of g indicate the temperature for which the 


heat of the liquid is obtained. 
If the number of revolutions or double strokes N are assumed, 


the displacement of the water end will be 


G Ibs. per hr. 


D3 = oN cubic feet. 


The air end of the pump is made in many cases of empirical 
design. Some authors give ratios of volume displaced by the 
pump per minute to the volume of the condensed steam or to the 
volume of the low-pressure cylinder of the engine which is dis- 
charging into the condenser. Several of these are mentioned. 


AIR AT PRESSURES BELOW THE ATMOSPHERE 29 


RATIO OF AIR CYLINDER DISPLACMENT TO LOW-PRESSURE CYLINDER 


Single-acting vertical pump surface condenser........ Ios 13 
single-acting ‘vertical pump jet condenser.:..-....-... 1: 9 
Double-acting horizontal pump surface condenser..... ae a 
Double-acting horizontal pump jet condenser......... be aes 
Double-acting horizontal pump-compound engine sur- 

PAGEECMTICLCLLSCLa writ Mer ase nak tag ao Tt 26 
Single-acting horizontal pump-compound engine sur- 

POE SECON COSCL etme Ee kaie (in adn Mee. hehe oe Read STO 


RATIO OF AIR CYLINDER De ee TO VOLUME OF CONDENSED 


SIIRIACOPCOLICL CHISEL Me Mine ret eee pe, tie ne Sn ee ee Te OG 
Net COD CDsct me a Ete Pac ok a heehee Mee oe Ee SAO 


This may be a satisfactory way, but it is better to estimate the 
volume from the air probably present. Water usually contains 
air to about one-fifteenth of its volume. This amount of air is 
at atmospheric pressure fa. and it must be cared for by the air 
pump at a reduced pressure. In addition to this there are small 
leaks in the pipe line which.allow more air to enter. A small hole 
will destroy the vacuum of the air pump. To find the volume 
of air per minute the following formula will be used, allowing 100 
per cent. for leakage. 


= Ins I iy thy : : 
V=(@X2.Xg) (<-) (bps) T, cubic feet per min. 
p = absolute pressure in the condenser, pounds per square inch. 


ps = vapor tension or absolute steam pressure corresponding to 7¢.! 
T. = absolute temperature in condenser. 
Ta = absolute temperature of atmosphere. 

This equation shows the importance of making ps as much less 
than p as possible. The terms p and ps; do not differ much, and 
by taking the mixture of air and vapor on its way to the air pump, 
through as cold a passage as possible, the term ps; is made smaller 
and the denominator is increased, making V small. This is the 
reason for the great advantage in a counter current for condensers, 
and even in the condenser, shown in Fig. 14, the coldest water should 
enter directly over the air-pump inlet so as to cool the mixture 
going to the pump. 

From the volume thus computed the displacement of the air 
pump is given by: 

Dar=sy cubic feet. 


1 See discussion on Partial Pressures Appendix C. 


30 AIR COMPRESSION AND TRANSMISSION 


Knowing the displacements of these pumps a stroke may be 
assumed, and from it the area determined. 


A es sq. ft. for the water pump. 


Aer=—¢ sq. ft. for the air pump. 
Steam Cylinder Size.—The cards from the water end are shown in 
the lower part of Fig. 15, while those for the air end are shown 


above. The combination of these or the addition of them when 


¢ Atrnosphere 

‘N 

n 
N 
N 
N 
N 
N 
N 
N 
N 
N 
N 
N 
N 
Ne 
N 
N 
N 

a 

7 S fe 

Atmosphere ‘ ie 
Ure 
Fic. 15.—Indicator cards of Fic. 16.—U. S. Navy pump 
condenser pump. cylinder. 


reduced to the proper scale, on account of the difference in piston 
area, will give the total work, and from this the area of the steam 
cylinder may be calculated, if the mean effective pressure 
(M.E'P.); be found fora given boiler pressure. Allowing 33 per 
cent. for friction, which is made large to give certain driving power, 
the following results: : 


4 LEP. )arAap+(M-EP.)pA p 
a iT.0O— 0.23) LUGE ta). 


Smit. 


U.S. Navy Air Pumps.—Separate air pumps are often used. Fig. 
16 shows the air cylinder of a stream-driven pump used in the U. S. 
Navy. This air pump is made with two air cylinders driven through 
gears from a steam cylinder placed on one side of a pump barrel. 


AIR AT PRESSURES BELOW THE ATMOSPHERE 31 


The pump is of the bucket type with foot valves AA and head valves 
at B. These with the valves in the bucket at C are all spring-con- 
trolled metal valves. The foot valves aie placed on an inclined 
partition for the purpose of making it easier to discharge the air when 
the piston rises and forms a vacuum. The lip around the discharge 
valve makes a dam and covers the valve with water. This makes 
them air tight. The other valves are also flooded, since all of the 
water on the bucket or that over the foot valves cannot be driven 
out, as the valves limit the motion of the bucket. On thedown stroke 
of the bucket the pressure in the space above it soon falls to a low 
vacuum because it had been completely filled with water; this, then, 
causes the valves to open and take air from the lower portion of the 
cylinder. ‘The air in the water also separates and rises to the top of 
the cylinder. Finally the bucket reaches the water below, and this is 
driven through the valve openings which are uncovered. It is seen 
that the air leaves first in this case. The water is struck by the 
bucket surface and will cause considerable shock if the pump is run- 
ning too rapidly. | 

Edwards Air Pump.—To do away with shock and to decrease 
valve resistance, the Edwards air pump, Fig. 17, was introduced. 
In this air pump water and air enter 
the space A at the bottom of the pump 
which is made conical in form. The 
piston B, which is driven from the 
steam piston by two rods CC extending — 
over the shaft and crank, is provided 
with a conical bottom. As this piston 
descends there is a vacuum produced, 
so that when the top of the piston 
uncovers the openings #, air enters 
{from the space A around the cylinder 
barrel, and as the conical bottom enters 
the water in the bottom of A, this is forced around the curved passage 
and discharged into the openings at HF. This continues even after the 
piston starts up, as the momentum causes the water to continue its 
motion. This discharge of water into the openings as the piston is 
moving upward acts as a valve to keep the air from coming out as the 
piston ascends. In a short time, however, the piston covers the 
ports or openings E and then the air and water are compressed 
until the pressure is sufficient to overcome the atmospheric pressure 
on the head valves at H, which are flooded by means of a lip around 


rN 


G 


NOS 


l 
NS 


<<) 
a 


Fic. 17.—Edwards air pump 
cylinder. 


O2 AIR COMPRESSION AND TRANSMISSION 


the valve deck. The piston rods CC are carried through long-sleeve 
stuffing-boxes so arranged that the point H, at which leakage could 
occur, is water sealed, leaving only one stuffing-box at the plate K to 
carefor. Thisis a simple matter. 

Industrial Uses of Vacuums.—One of the earliest applications of 
air pressure below the atmosphere is shown in a patent dated 1833, 
for the preparation of leather by the evaporation of certain substances 
in a partial vacuum, the object being to avoid intense heat. Water 
at atmospheric pressure boils at 212° F. If the pressure is increased 
above the atmosphere as in the ordinary boiler, the boiling-point 
of the water is raised, and for the same reason when it is desired to 
evaporate any substance at a temperature below its boiling-point 
for atmospheric pressure it is merely necessary to put the substances 
in a partial vacuum and its boiling-point is accordingly lowered. 
This results in evaporation at very 
low temperatures, a most desir- 
able feature especially in the drying 
of fruits, etc. 

Air at pressures below the at- 
mosphere is used for drying all 
kinds of food materials such as 
meat, fish, fruits, etc. Frequently 
a solution of a gelatin sugar or 
gum is used as a coating. 

Vacuum processes are employed 
for pickling and salting meats and 
vegetables, evaporating fruits, re- 
fining sugar and condensing milk. 
Wood may be artificially colored and railway timbers impregnated 
with preserving substances by means of a vacuum. 

Salt Evaporating Effects.—Probably one of the most interesting 
applications of the partial vacuum is found in the manufacture of 
salt, the refining of sugar and the concentration of syrups, liquors, 
etc., by what is known as the triple effect apparatus. 

The apparatus in which this is done usually consists of two or 
three ‘‘effects,’’ as they are called, almost identical in construction. 
One of these is shown by Fig. 18 and represents one of a “‘train”’ 
of effects for extracting salt from brine by heating the brine solution, 
thereby evaporating the water and leaving the salt as a solid deposit. 

B is the heating chamber or section consisting of a series of vertical 
flues, conical in section, in which the brine circulates and around 


Ei 


Fic. 18.—Vacuum manufacture of salt. 


ALK AIP PRESSURES BELOW THE ATMOSPHERE 33 


which the steam flows. This part is very similar in construction to 
a vertical flue boiler, with the exception that the flues are conical 
instead of cylindrical, to prevent deposits on the tubes. 

Steam js furnished to B through pipe £ either from a boiler or 
the exhaust of a steam engine, and after giving heat to the brine, 
which fills the apparatus as shown, the steam is condensed and 
drawn off. ‘The vapor from the brine in the first effect passes through 
pipe / into the heating section of the second effect, gives heat to 
brine in the second effect and in doing so is condensed. This con- 
densation of the vapor produces a partial vacuum in the first effect, 
thus lowering the boiling point of the brine in that effect and hence 
aiding evaporation. 

The vapor of the brine of the second effect is conducted to the 
heating chamber of the third effect, imparts heat to the brine in that 
effect, producing a partial vacuum as in the first instance. The 
vapor from the third effect passes to an air pump. ‘This air pump 
maintains a good vacuum in the third effect and in consequence 
the boiling-point of the brine in that effect is very low indeed. 

The vacuum in the second effect is not good as in the third, and the 
boiling-point of the brine in that effect is a little higher. In the 
first effect the poorest vacuum exists and in consequence the brine 
here has the highest boiling-point of all and therefore requires the 
most heat, which is supplied to it from the boiler or the exhaust of 
the engine operating the air pump. It is because of the different 
boiling-points of the brine in three effects that the heat of the vapor of 
brine in the first effect is enabled to evaporate the brine in the 
second effect and the vapor of brine in the second effect is able 
to evaporate the brine in the third effect. In each one of the three 
effects of this apparatus salt is being extracted, the solid matter 
settling to the bottom of the chamber C from which it can be with- 
drawn by means of the two valves without disturbing the operation. 
The salt and some brine with it are deposited on a filter in the cham- 
ber D. The salt is here washed and the brine below the filter is 
returned to the evaporating chamber through H. The salt is then 
removed and a new supply of brine introduced through G. In 
this way the operations can be made practically continuous and 
in many plants automatic. 

In some of the salt machines, instead of having two valves for 
withdrawing the salt without disturbing the partial vacuum, the 
lower part of the apparatus consists of a pipe running down such a 


distance that the partial vacuum will equal the hydrostatic head 
3 


o4 AIR COMPRESSION AND TRANSMISSION 


and consequently the salt may be removed without the use of any 
valves whatever. 

Concentration of Liquors.—In the concentration of syrups and of 
many liquors, it is highly important that the liquors should not be 
heated too highly or they will be scorched. To avoid this, the liquor 
is moved through pipes at a velocity so great that there is no oppor- 
tunity for the syrup to become scorched. 

In the concentration of liquors, a partial evaporation is secured in 
one effect and the vapor of the liquor separated from the liquor | 
itself. Both vapor and liquor are then introduced into a second 
effect which has a lower pressure, and here the vapor from the first 


o—> 
< 


aELer 
=a) 
ze 
2) 
2 


F| | 


= 


Fic. 19.—Vacuum concentration of liquors. 


effect gives heat to its own liquor and the liquor is still further 
concentrated; the resulting liquor and vapor of this second effect 
are separated as in the first instance and introduced into a third 
effect where the concentration is carried still further. In some 
evaporators this is continued in a fourth effect and a still further 
concentration of the liquor secured. 

Figure 19 shows an evaporator of the type just explained. 
The operation is as follows: The steam which may be either the 
exhaust from an engine or live steam from a boiler is led into the 
cylindrical chamber through A. The liquid to be concentrated is 
fed in through the tubes B and enters the evaporator in a small 
but continuous stream and immediately begins to boil violently, 
becoming a mass of spray containing, as it rushes along the heated 
tube, an increasing proportion of steam. The outlet of the tube C 
being at a lower pressure than B, the contents are propelled through 
the tubes at a high velocity, finally escaping into the separator D. 
Here the steam or vapor with its entrained liquid is discharged with 
considerable force against the baffle plates H, causing the liquid to 
be separated from the vapor, the concentrated liquor being drawn 


AIR AT PRESSURES BELOW THE ATMOSPHERE 35 


off through a trap F, while the vapor escapes through G to enter the 
second effect where its heat still further concentrates the liquid, 
which is conducted from F of the first effect to the second effect, 
entering through pipes similar to B. The liquid is led from the bot- 
tom of the separator of the first effect into the coils of the second 
effect and is further concentrated, passing in this way through the 
entire system or “train.” 

The volume of the liquid is being continually reduced as it passes 
through these effects, and as the pressure falls in passing from one 
effect to the next the boiling-point is lowered. That in the last 
effect being the lowest of all, the required low pressure for this effect 
is secured by a vacuum pump. This relative reduction in pressure 
and consequently of the boiling temperature automatically adjusts 
itself, no matter how many effects are used, thus effecting the boiling 
of the liquid by the steam produced from the same liquid when in the 
preceding ‘‘effect.”’ 

One of the advantages claimed for this system of evaporation of a 
liquid in the form of a spray subjected 1 to heat under a vacuum is 
that it receives the heat 
quickly and is concentrated 
slightly at one temperature, 
then still more at the lower 
temperature of ~ the: next 
effect, and so on, thus re- 
ducing the danger of over- 
heating. The rapid move- 


ment of the liquid aided by =, ; 
the vapor which is moving a 
in the same direction keeps ) 


the liquid in the form of a 

spray, thus taking up very 

quickly the heat given to it. i OTT, 

Evaporation of Cane Juice. pre. 2o,—Vacuum concentration of liquors. 

—The cross-section of a still 

different type of evaporator is shown in Fig. 20. The liquid or cane 
juice is introduced through A and is sprayed from holes in pipe B over 
a series of steam-pipes. The partially concentrated liquor falls into 
chamber C and is drawn from there by a centrifugal pump D and 
forced into the next effect through a pipe similar to A and B of the 
first effect. Steam for heating the first effect is introduced through 
pipe E and, after imparting heat to the liquor, is condensed. The 


36 AIR COMPRESSION AND TRANSMISSION 


water falling to the bottom, as shown, is drawn off through F. The 
vapor from the liquor that has been partially concentrated escapes 
through G and is introduced to the heating chamber of the second 
effect through a pipe corresponding to £ of the first effect. As the 
pressure of the second effect is lower than the pressure of the first 
effect, the condensed steam from this effect is also introduced into 
the second effect, and, being at a temperature above that of boiling- 
point of water at this lower temperature, it gives heat to this effect 
and helps to evaporate the liquor in it. 

Vacuum Cleaners.—One of the most recent applications of air 
at pressure below the atmosphere consists of vacuum cleaners of 
various types for removing dust and dirt fiom floors and walls, 
furniture, etc., in buildings. Although many types of machines 
for producing the 1equired vacuum are on the market, they may be 
grouped under two heads, namely, portable and stationary types. 
In the former type of vacuum cleaner, the machine is moved about 
the room, drawing dust and air through a pump and discharging 


Fic. 21.—Syphon. 


into a cloth receptacle, from which the air can escape and in which the 
dust is trapped. This type of cleaner has the advantage that it is 
comparatively cheap, but has a disadvantage in that germs are 
discharged with the air into the room, and from the hygienic point 
of view this is considered objectionable. 

The objection mentioned is removed in the second type of cleaner, 
in which the machine is permanently located in the basement of 
the building and vacuum pipes lead to the various rooms and floors, 
to which the hose and cleaning tools are attached. This type of 
machine is naturally more expensive, but finds considerable favor in 
large office and hotel buildings. 

The piston type of air pump is used as well as fans, bellows, and 


AIR AT PRESSURES BELOW THE ATMOSPHERE 37 


rotary blowers. This field of usefulness for air at low pressures has 
increased to a remarkable extent, but the industry cannot yet be 
said to be on a fixed basis; that is, there is still considerable informa- 
tion needed regarding the proper suction pressures to secure the req- 
uisite cleanliness without injuring the fabric of rugs: and hangings. 

Syphon.—A discussion of the uses of air at pressures below the at- 
mosphere would not be complete without some reference to the 
syphon (Fig. 21). Bis an air chamber, C a water seal for the valve F, 
D a funnel for filling the syphon and also for sealing valve K against 
air leakage. After the syphon has been filled, valve K is closed and 
G and H opened. This starts the syphon in operation. The air 
that comes in with the water and through the joints of the pipe 
collects in chamber B and may be discharged by closing valve F, 
opening valve K, and filling chamber B with water. Then close 
valve K and open valve F and any air below C will rise into chamber 
B and the water will take its place without stopping the running 
of the syphon. 


CHAPTER VI 


AIR AT LOW PRESSURES 


Uses of Air at Low Pressures.—Probably the principal uses made 
of air at low pressure are for cupolas, furnaces, blacksmith fires, for 
conveying light materials, as shavings, etc. Avery large field for air 
at low pressure is for purposes of ventilation for school buildings 
churches, theaters, assembly halls, factories, mines and tunnels, etc. 
Its use in this field dates back to the sixteenth century, and while for 
many years very little thought was paid to it, to-day considerable 
attention is paid to the subject of ventilation. In fact, no heating 
system for home, school-house, factory or office building is complete 
without some system for removing the foul air and replacing it with 
fresh air. 

To secure the necessary movement of air in buildings where the 
number of cubic feet of room per person is a limited quantity, a 
positive circulation is secured by introducing the fresh air at a pres- 
sure a few ounces above the atmosphere into the room, or by drawing 
the foul air from the room by means of an exhaust fan. 

Compressors for Low Pressures.—The principal machines used 
for moving air for ventilation and other purposes, either by pressure 
or suction, are: the centrifugal fans or blowers, the positive blower 
of the piston or rotary type, and the jet pumps from which are dis- 
charged jets of steam or compressed air. The requirements for good 
ventilation demand that large volumes of air must be moved at com- 
paratively low velocity and pressure, which is not a favorable con- 
dition for high efficiency and can in general be better satisfied by a 
centrifugal fan or blower than by any other machine; it may also 
be stated that the fan is comparatively cheap to install, is simple in 
construction and possesses a fair efficiency. 

Tables giving requisite information regarding fans for various 
purposes can be secured from any of the fan manufacturers, and 
engineering handbooks usually contain considerable data taken 
largely from these catalogs. 

38 


AIR AT LOW PRESSURES 39 


It is well to remember, however, that these tables are apt to over- 
rate the capacity and under-rate the required power for operation. 
Centrifugal fans are in use furnishing air at pressures varying from 
1/4 0z. to 20 oz., and are constructed in all sizes, the largest, of course, 
being used where large volumes of air are to be moved at very low 
velocity. 

. Air for Forges.—An article by William Sangster in the Transac- 
tions of the American Society of Mechanical Engineers, in Volume 22, 
page 354, gives the following approximate rules of the air required 
for forges and cupolas. The maximum pressure required for forges 
is about 4 oz. per square inch, the ordinary pressure about 2 0z.; 
140 cu. ft. of free air per minute is ample and it is estimated that it 
requires about 1/4 h.p. to furnish air for an ordinary forge. It is 
customary to estimate that an exhaust fan for a blacksmith-shop 
must remove four times the amount of air delivered at a pressure of 
3/4 0z., and that to do this will require about 1/5 h.p. per forge. 
Roughly speaking, if the number of forges is divided by 4, the horse- 
power required to furnish the blast can be found, and if the number 
of forges is divided by 5, the horse-power required to exhaust the 
smoke can be found. 

These exhaust fans run at a much slower speed than the pressure 
fans and as the pressure of the exhaust air is much lower than the 
blast the power required for their operation is less, although they 
move four times the volume that the pressure fans do. In some 
installations one fan does the work of forcing the blast and ex- 
hausting the smoke, but as the requirements of a blast fan are so 
different from those of an exhaust fan, such a combination is not 
economical. 

Air for Cupolas.—The air required to melt iron in cupolas may be 
taken as 40,000 cu. ft. per ton of iron melted, and the horse-power 
required as three-tenths of the number of tons to be melted per hour, 
multiplied by pressure of the blast in ounces per square inch. It is 
well to remember that these figures do not take account of losses in 
the piping system. These results will, no doubt, fall far short if 
the pipe system is poorly designed with sharp elbows, small diame- 
LErsy tc: 

Air for Ventilation.—In estimating air required for ventilation, the 
data in Table IV is frequently used: 


40 AIR COMPRESSION AND TRANSMISSION 


TABLE IV.—AMOUNT OF AIR REQUIRED FOR VENTILATION 


Allowable parts of carbonic Cubic feet of air required per person 
acid in 10,000 of air 
Gr COR Per minute | Per hour 
5 100 6,000 
6 50° 3,000 
7 33 2,000 
8 25 1,500 
9 20 1,200 
IO 16 1,000 


TABLE V.—AIR SUPPLY FOR VARIOUS BUILDINGS 


Cubic feet | Cubic feet 


Air supply per occupant for Pe Aras Heche 
Hospitals.2.toy chess Get oo ae le he ae ONTO TLOD. 4,800 to 6,000 
High schools. . Le ee ate Oe Brae 50 3,000 
Grtcrcaccis® te erty, hc BS en ln ce 40 2,400 
Theaters and aesenibly nail eae Wears: De 1,500 
Ghrarchee an eee ak a ae 20 1,200 


TABLE VI.—AIR SUPPLY FOR VARIOUS ROOMS 


Use choos Changes of air 

per hour 
Public.,waiting-room /.. 1.621, dee ol eee ee 4 to 5 
Publicstoilets (ces 85h el eee 5 to 6 
Coatvand Hocker-rooms.”.. 2... 4. ee eee ee 4-to 5 
MMIUSOUIIS BS ence cite «Sch plc athe EEA oe eee rena 3to4 
Offices;*piubltes: his iol a, see 4 to 5 
Offices private < asl. cst. c Se ee ee ee ee 3 to 4 
Public: dining-rooms, 5395 eee ee 4to5 
Living-ToOMS WR eee cn Sao een eee pnd Ne eee 3 to 4 
Libraries, public. . 5 safle: Bieta d seiaced, ewe TERee ee tues aa eee mT 4 to 5 
Libraries, is ae ch SS ee ince Ee cae 3to4 
Fuming cabinets fe eivctaica jaboratorres Lo a AOL eh one we ee ree 30 to 60 


The following material on fans and blowers is taken from a lecture 
by Mr. H. de B. Parsons, Consulting Engineer, delivered before the 
Junior Class of Columbia University: 


LECTURE BY MR. H. DE B. PARSONS 4] 


FANS OR BLOWERS 


‘A fan or blower is a machine for impelling gas, z.e., for producing 
a current of gas. In the majority of cases the gas impelled is air. 

“There are many purposes for which a fan is used, such as for 
heating, cooling and ventilating buildings, either by exhausting air 
from, or forcing air into, the apartments; for blowing the fire of a 
forge or cupola; for creating artificial draft for fuel combustion; for 
work pertaining to drying; for carrying away obnoxious gases and 
discharging them at a point where they will not create a nuisance; 
for carrying away grindings and waste products so that they may 
not affect the workmen; for conveying light materials, such as saw- 
dust and small particles, and permitting them to settle in dust 
chambers; and for the circulation of air in mines and places where 
explosive gases may collect. 

“When fans are properly selected for their work they will give 
satisfactory and economic results, and will require little attention 
for maintenance. 

“The conditions of pressure and density of the gas and of speed 
and capacity of the fan govern the size, type and proportion of the 
fan and its housing. These conditions are closely related, and all 
affect the design that should be selected. Even moderate differ- 
ences in the conditions of operation will have considerable effect 
upon the power necessary to drive the fan. It therefore follows 
that a fan should be designed for the conditions under which it is 
to operate, and conversely, that a fan should be operated under the 
conditions for which it was designed. 

‘Fans are not economical machines to operate against high pres- 
sures. In such cases a blowing engine or compressor will be the 
Detter 


Classification 


“There are a number of types in use, but nearly all blowers and 
fans can be classified under one of the follcwing heads: 
(1) Rotary blowing machine. 
(2) Disc, axial or propeller wheel fan. 
(3) Centrifugal fan, either a fan blast or cone wheel. 
(4) Turbine blast or high-speed centrifugal fan. 
“Type (1) is a positive or displacement discharge machine, and 
is a blower or exhauster. 
“Type (2) is an axial discharge fan. 
“Type (3) and (4) are peripheral fans. 
‘All of the types can be used for exhausting or for blowing, 
although some are less suitable for exhausting than others. There 
is a material difference in the selection of a type for an exhaust 


42 AIR COMPRESSION AND TRANSMISSION 


machine, when pressures above the atmosphere on the discharge 
side of the fan are considered. The disc fan makes a gcod exhauster 
when a pressure above that of the atmosphere does not have to be 
maintained on the discharge side, but when such a positive pressure 
has to be maintained a centrifugal machine is the more suitable. 


Definitions 


“There are certain terms used in fan work which are recognized 
as having specific meanings. 

‘Fan Pressure or Draft——Fan pressure or draft means the 
difference between the pressures on the suction side and on the 
discharge side of a fan. The difference in pressure is expressed 
either in ounces per square inch or inches of water. 

“When a fan is used as an exhauster discharging into the atmos- 
phere, there will be a partial vacuum on the suction side and slight 
pressure on the discharge side. In this case the vacuum is expressed 
as the number of ounces or inches of water below the atmosphete, 
and the fan pressure or draft is measured by the difference. It is 
just the same as if the suction were at atmospheric pressure and the 
discharge at the same number of ounces above the atmosphere. 

‘Fan Capacity.—Capacity means the maximum discharge of free 
air froma fanin cubic feet per minute against a pressure cor- 
responding to the speed of the tips of the blades. This condition is 
satisfied in the case of a centrifugal machine when the velocity of 
the gas entering the inlet is equal to the velocity of the inner edge 
of the floats at inlet. 

‘““Housing.—The casing in which a fan operates is called the 
‘housing.’ It is made of metal, of brick, or of wood. Frequently 
the fan is so set as to project into its foundation, and in such cases 
the casing only covers the portion which projects above the founda- 
tion. ‘The fan is then said to have a three-quarter housing. Of 
course the inlet must be above the foundation or a free passage must 
be provided to it. 

“Free Discharge.—A fan is said to have free discharge when the 
blast is free or unrestricted. This condition is maintained when the 
total head is practically equivalent to the velocity head. The total 
head is equal to the velocity head plus the friction head, and with 
a free discharge head the friction head is practically zero. 

“Restricted Discharge.—A fan is said to have a restricted dis- 
charge when the blast is restricted by ducts or by pressure reservoirs. 

“Free and Restricted Suction.—Similarly to free and restricted 
discharge, a fan may have either a free or a restricted suction, 
whereby the gas has either a free or unrestricted entrance into the fan, 
or a restricted entrance caused by ducts or a reduction in the pres- 
sure on the suction side. 


LECTURE BY MR. H. DE B. PARSONS 43 


“Coefficient of Contraction.—The ratio of the area of the vena 
contracta to the area of the orifice is called the ‘coefficient of con- 
traction.’ 

‘Coefficient of Velocity.—As the stream of gas passes the vena 
contracta its velocity is somewhat increased, and the ratio of the 
actual velocity to the theoretical velocity is called the ‘coefficient 
of velocity.’ In a well-shaped delivery orifice this coefficient of 
-velocity is not far from unity. 

“Coefficient of Efflux.—The ‘coefficient of efflux’ is the product 
of the coefficient of contraction and the coefficient of velocity. 

“Volume of Discharge.—The volume of gas discharged by a fan 
is a function of the product of the velocity of the gas times the area 
of the outlet, times the coefficient of efflux. 

“Blast Area.—The blast area of a fan is the theoretical area of 
outlet whose coefficient of efflux is unity. The volume of discharge 
is equal to the blast area times the velocity of discharge. Therefore, 
the blast area equals the capacity divided by the velocity due to the 
velocity head. The stream of gas issuing through an outlet is re- 
duced in area, depending on the shape and character of the orifice. 
This reduced area is called the ‘vena contracta’ and is usually at a 
distance from the opening of about half its diameter. This vena 
contracta is caused by the change in direction of the flow of the mole- 
cules of the gas as they pass the opening.”’ 


Measurement of Draft 


‘The measurement of draft, either static or both static and veloc- 
ity pressures is obtained by noting the difference in level of a liquid 
in the arms of a tube bent on the form of a U, of which one end is 
connected with a proper tube to the space in which the draft is to 
be measured. The liquid is generally water although for heavy 
pressures mercury is sometimes used. 

“There are different forms of gages which can be bought in the 
market but for ordinary work the simplest forms are the best. 
Some of these instruments are so made that they will give a con- 
tinuous record, and for certain kinds of work these continuous 
records are of considerable value. 

‘““Anemometers are used for measuring the velocity of the gas. 
Readings should be taken at many points in the cross-section of the 
current, and even at the same points consecutive readings will not 
agree. Multiple readings therefore should be made in order to aver- 
age up these irregularities. 

“Fan draft is always expressed in ounces per square inch or in 
inches of a water column whose weight is equal to the ounces per 
square inch. The velocity corresponds to this pressure, when the 
friction head is zero. 


44 


AIR COMPRESSION AND TRANSMISSION 


‘When the pressure exceeds two or three pounds per square inch 
as is the case with many positive blowers, the pressure is then 
generally expressed in ‘pounds per square inch.’ 


Corresponding to Various Heads of Water in Inches 


TABLE VII.—PRESSURES IN OUNCES PER SQUARE INCH 


Decimal parts of an inch 


Head 
in 
inches 
0.00) 0.8 ome O28 0.4 On5 0.6 OF 0.8 0.9 
Osmaioe ae 0.00510, 12] O78 7 0,23 OF 209) Ones alo. 10 sO ntOm Oma 
I 0.58 |)0703° 190.00 | Oa7s 470 Siw Ous7 IFO..03 4207 O0om tie O4mmimoG 
2 EALO WAL Rea bay a Fee eT ee) MAT aTR COM Tec Op) at Ot ee 
3 Te73UESFO |) Te BS |ATOL | 1.06) 2.02 1e2. OS 2b TA hee Om meee 
4 2SAT VW 2e27 N24 2 I 2048 PO RAT eaAOO NO Oe OO aD. 72 mire ae james 
5 2280 1-2.04 11-3500) | 3200) - 297 2elee rio 47 24 mg 20m) ec ns 5 amr ey 
6 P3247 13-521 3258 |, 3404 53 FOS <7 6 Woh Olas ee 7a) ns cOe mt ReaD 
7 AVOAY 4.10 144.10 4322 WAI 28 CA ea AO ed Calafate ee 
89) 4024 4007 A273 104.90 W485 des Ot 4.07 15.03 195 Om es ae 
9 5.20) 6:26 | 8.35 | a7 te Ae US lace sd ee se O081 SOO mms age 
TABLE VIII.—HEIGHT OF WATER COLUMN IN INCHES 
Corresponding to Various Pressures in Ounces perc Square Inch 
Eres: Decimal parts of an ounce 
sure 
in oz 
per 
: ‘ : : ; é .6 : 8 O: 
rae 0.0 Cut One 0.3 0.4 O25 fe) On ° 9 
fe) Sf) OnE 1) O35.) -Ol'5 2 a2 0G LOT IAA OAM Walwo2 Taler sot ate et 
I £272) (1.00) -2.08:) 92, 00 12. Al) 22800 aan 7 7 Od abet ke 
2 3. AON 935631053. 81 }) 3.08 ATS (ASS etn5O eed. OF sd O4 sees mon 
3 ST LOW eS Se Gal ee a7 EMS oe 106.1. 6503 1 Gi40 8102571 Oars 
4 6.02) IvOON 727") 72 AA FROEN Aeon E00 sO aes terol sO MO wAS 
5 8.65]. 8782 |G. 00," 0.17) 02344 0.5218 On00 sO. 004510503) Ione. 
6 10.38 | 10. $5 (10.73 [TO-DO MIT O7 UT En20 hin A set COG tis 7 Fer reo 
7 12.1} 12,28) @o. 40142 363 12, SON Os Lowel os 2 el oi Ou Toe 
8 13.844 TA, 00 | T4479. 014. 260 14a Sa 14s 71 ae Sow 5 Obe LS. Sea 
Oh 18.871 15 574 hFS) 02) 16-00 i1G220 10 48 nOuOe W 10. 70) 20.00) fy 


DCECTORE BY MR. H. DE BY PARSONS 45 


Fan Efficiency 


“In the operation of a fan or blower there are certain losses 
utes must exist, the principal losses being: 

‘rt. Fluid friction and eddies caused by the movement of the 
gases. 

‘““9, Leakage of the gases backward through the fan or blower. 
This is sometimes called the ‘slip.’ 

“2. Mechanical friction of the moving parts of the apparatus. 

‘“‘ Generally speaking, these losses increase with the speed of revolu- 
tion of the fan and also as the difference in pressure between the 
suction and discharge sides increases. 

haiine efficiency of a fan or blower is the ratio of the useful work 
done on the air divided by the work required to drive the fan. Fans 
are generally driven by steam engines or by motors, and frequently 
the denominator of the efficiency ratio includes the work of driving 
the engine or motor. Such an efficiency is really the combined effi- 
ciency of the prime mover and fan. 

“The efficiency of a fan wheel with a housing varies with the ratio 
of diameter of inlet to diameter of wheel. The smaller this ratio 
the greater will be the theoretical efficiency so long as the area of 
outlet times the coefficient of efflux is not less than the blast area. 

‘““When a centrifugal fan has to work against high pressures, it is 
desirable, therefore, that the ratio of inlet to wheel diameter be small 
in order to get the benefit of this increase in efficiency.”’ 


Flow of Gas Through an Orifice 


‘““Gas flowing through an orifice does not obey the same law as 
the flow of fluids. The reason of this is that gas expands from 
the higher pressure to the lower pressure as it issues through the 
orifice. 

‘“‘Tmagine the gas in a reservoir R (Fig. 22) flowing from the short 
cylindrical orifice of section a. Imagine that the reservoir is kept 
supplied with gas so that its pressure remains constant. Suppose 
that the division S represents a pound of gas. 

‘““As the gas escapes through the orifice a, the pressure is kept 
constant, and the work OA EQ has been done upon the gas. The 
gas in expanding develops the expansive work EIMQ, EI being an 
adiabatic curve. 

“The outer pressure P, absorbs the work JVOM, and the balance, 
AEIN, is devoted to accelerating the particles of the pound of gas 
to a velocity v. 


46 AIR COMPRESSION AND TRANSMISSION 


“Hence the work AEIJN equals the actual energy of 1 lb. of gas 
moving with a velocity of v ft. per second or 


a 
AEIN =(t lb.) X — 
2g 


Therefore v=/2g (AEIN) ft. per second 
Taking the law of the curve FI as p101)"=fov2"=constant we 
have: 


fi 
nN pe a 
AEIN=""_ xX 144 piS ae (2) | 


1 


Fic. 22.—Flow of gas through an orifice. 


“Letting m =1.405 the ratio of the specific heats and the exponent 
in adiabatic changes of air 


‘ihe a [28x25 X 144 p:S|1— @) be 


v= 1794 | P15 E —~ C) oe) ft. per second 


“Tt is found by taking the pressure at the orifice by a gauge G that 
if the gas flows into another reservoir kept at the back pressure py» 
the orifice pressure is identical with 2 if the latter is more than 
about 0.581. That is, if p1 is r00 lb. per square inch absolute, any 
value of the back pressure greater than about 58 lb. gives this 
pressure to the orifice, but if the back pressure is the atmosphere 
the orifice pressure remains 58 lb. 


“Tt therefore follows that for ao 58 the velocity given by the 
1 
above formula exists at some point of the jet beyond the orifice at 


LECTURE BY MR. H. DE B. PARSONS 47 


a section a,> a due to the natural spread of the jet, while the velocity 
at the orifice or throat of the jet is that given by the formula for 


Loss of Head Due to Friction in Ducts 


“The frictional resistance to the movement of a gas in a duct is 
proportional to the surface of the duct. It is, therefore, directly 
proportional to length and inversely proportional to diameter. It 
also varies as the square of the velocity. 

“Therefore, the ducts should be of ample area, or the power lost 
in friction will be very great. Small pipes and high velocities 
should be avoided. 

“Tt is evident that after a certain size of duct is reached, any 
further change in size or velocity of movement will only have a rel- 
atively small effect upon friction loss. The limit, therefore, is 
reached when the increase in space required and the cost will turn the 
saving in friction into a loss from a commercial standpoint. 

“Usual Velocity in Ducts.—In heating and ventilating work for 
theaters, hospitals, churches and large buildings, the limiting 
velocities usually selected are: 


(a) In ducts leading from force fans— 


In borizontalimaineducts 27 427. 2-1 ,000 It, per minute 
imshorizontalimatnibranches #42. 72.428... a: 1,300 ft, per minute 
In horizontal branches to risers............. 650 ft. per minute 
tie WerticalsTisers eau woe es Sin ison Sa bes Ae 800 ft. per minute 


(b) In ducts leading to exhaust fans— 
EWevertical riscrs ea erty or aan Goo 1b per minute 
Tnehorizonraduets tolans 2 sane tears £000 ttn per minute 


‘“The frictional loss in ducts can be calculated from the formule 
for the movement of fluids. In addition to the friction loss of head 
caused by the passage of a gas through a straight duct, there is a 
loss at each bend or change of section. 

“In order to overcome these friction losses, it is necessary that the 
pressure at the fan end of the duct should equal the sum of the 
pressure desired at the open end of the duct, and the pressure 
necessary to overcome the losses in frictional head. 

“Tn all the following formule the following notation has been used. 


Notation of Symbols 


A denotes the area of the duct in square inches. 
a denotes the blast area, or the ‘effective area of discharge’ in 
square inches. 


48 AIR COMPRESSION AND TRANSMISSION 


denotes the diameter of duct in inches. 

denotes the perimeter of duct in inches. 

denotes a constant. 

denotes the diameter of the fan wheel in inches. 

denotes the density of the gas, 7.e., its weight in pounds 

DerecuLaLe. 

E denotes the combined efficiency of fan and its prime mover. 

e denotes fan efficiency, or ratio of useful work to work of driving 
fan. 

g denotes the acceleration due to gravity in feet at the end of one 
second, 32.16 ft. 

h_ denotes the equivalent head, 7.e., the height of a column of gas 
in feet having a density d, Sihaee mene will produce the velocity 
pressure p ounces per square inch. 

K denotes the capacity of a fan in cubic feet per minute. 

1 denotes the length of duct in feet. 

nm denotes the number of revolutions per minute of fan wheel. 

Ie 

Pp 


Stef ope. 


denotes the total pressure against which a fan is working.! 
denotes the velocity pressure in ounces per square inch (or inch 
of water) against which a fan is working. 
t denotes the absolute temperature, or 460°+ F. 
Q denotes the volume of gas discharged by a fan in cubic feet per 
second. 
V_ denotes the peripheral velocity of fan wheel in feet per second. 
vy denotes the velocity of gases in feet per second due to pressure p. 
W denotes the width of fan wheel in inches. 
w denotes the width of blades of fan wheel at periphery in inches. 
‘Pipe Losses.—Frictional losses are very hard to calculate, as 
so much depends on the smoothness of the surface and the material 
of which the ducts are made. 
“The loss due to surface friction can be estimated from the 
formule: 
“For circular ducts of galvanized iron, carefully made, 


“For rectangular ducts of galvanized iron, carefully made, 
ly?C 


100,000 A 


in which p denotes the loss of pressure in ounces per square inch. 
This is an empirical formula based on Weisbach’s general formula 
for the flow of fluids. 


1 The total pressure against which the fan is working is +s, in which ps is 
the static pressure. 


LECTURE BY MR. H. DE B. PARSONS 49 


‘“‘Bends create an additional loss which are hard to estimate. For 
all practical purposes the frictional loss due to bends can be estimated 
sufficiently accurately as follows, when the ducts are of galvanized 
iron, carefully made and of fairly smooth surface: 

(a) For right-angle bends with the radius at the root of the bend 
equal to one duct diameter, allow an equivalent length of 
straight pipe equal to 
11.1 times the diameter 
Of the. duct.. [hus in 
Pig2923, 11 B =20-1n., 
allow for the bend 11.1 
X20 0r 222 in. or roft. of N 
pipe: : 

(b) For right-angle bends 
with a radius at the root 
of the bend equal to one- 
half the duct diameter, 
allow an equivalent 
length of straight duct equal to 29.5 times the diameter of 
the duct. 

(c) For 45-degree bends allow one-third of the loss for right- 
angle bends.” 


Fic. 23.—Right angle bend resistance. 


Rotary Blowing Machines 


‘“‘A rotary blower is a positive pressure blower or exhauster, and 
is not a fan, although it is used for similar purposes. It is positive 
in its action and it operates by displacement. 

‘A rotary blower costs more than a fan of equal capacity, but it is 
more economical than a fan when operating against high pressures, 
that is 8 oz. per square inch or more. Turbine blowers, however, 
are now being built giving efficiencies fully equal to that of rotary 
blowers. (See Chapter XI). 

‘A rotary blower is more economical than a compressor when 
operating against pressures less than 7 lb. per square inch. Gener 
ally speaking, the compressor is more economical at pressures In ex- 
cess of 7 lb. bn Lie 

‘‘Rotary blowers can be arranged to give constant pressures or 
constant volumes. They can also handle liquids as well as gases. 

‘““A rotary blower (Fig. 24) consists of a casing in which two 
impellers revolve in. opposite directions. Each impeller is of a 
double-lobe section symmetrical with its shaft. The impellers are 
set so that the lobe of one impeller fits into the recess of the other. 


The impellers do not touch each other, nor do they touch'the casing, 
4 


50 AIR COMPRESSION AND TRANSMISSION 


although they should work as close as is possible without touching 
so as to prevent loss through leakage. 

‘The air is drawn in through the inlet, is caught between the lobe 
of an impeller and the casing and forced around as the impeller 
revolves, and discharged through an opening situated in the casing 
diametrically opposite to the inlet. In order to keep the impellers 
at their proper relative speeds, one shaft is driven by the other shaft 
through a pair of gears. 


+ aa e+ CEPUCH DG a rn Tere 
I I< - Pitch Diameter. >| 


¢--~-- pitch Dia, -—---—-9| 


Fic. 24.—Cross-section through standard blower. 


“The pitch diameter of these gears controls the size and capacity 
of the machine. The radius of an impeller, or its half length, is 
made three-quarters of the pitch diameter of the gears. The casing 
consists of two semi-cylinders separated by a parallel section. The 
radius of the cylinders is equal to that of the impellers plus clearance. 
The width of the parallel section is equal to the pitch diameter of the 
gears plus the clearance. The speed of revolution is regulated by 
the safe speed at which the gears can be operated. 

‘“‘Blower Pressures and Capacities.—The limit of the gas pressure, 
in commercial sizes, is about 12 lb. per square inch. The standard 
commercial sizes have capacities varying from one-quarter of a cubic 
foot to 400 cu. ft. per revolution. | 

“Two of the types of rotary blowers in use, are described by the 
shapes of the ends of the impellers, as cycloidal or involute. 
When the impeller ends are cycloidal they fit close to each other and 
leave no waste spaces or pockets. Such machines are adapted to 
handle wet gases and liquids as well as dry gases. When the 
impellers are cycloidal, the capacity per each revolution is equal to 


LBC LU REeDYe Vik He DEO APARKSON S ol 


the area of the pitch circle of the gears times the length of the 
cylinder. 

‘When the impellers are involute, the capacity is somewhat greater 
than the cycloidal and depends on the diameter of the generating 
circle for the involute. This diameter is variable to suit the duty of 
the blower. 

“The slip is largest in small machines, and least in large ones. 
Thus for machines displacing three-quarters of a cubic foot per revo- 
lution at 1-lb. pressure the slip is about 60 to 70 revolutions, 7.e., 
the machine has to make that number of revolutions to hold the 
pressure against leakage. For machines displacing 300 cu. ft. per 
revolution at 1-lb. pressure, the slip is from 3 to 5 revolutions. The 
slip for intermediate sizes is about proportional and for pressures 
other than 1 lb. the slip will vary closely as the square root of the 
pressures. 

“For cycloidal types, the casing is 1 1/2-pitch diameters high by 
2 1/2-pitch diameters wide. For involute types, the casing section 
is nearly the same, but depends on the circle on which the involute 
is rolled, and this depends on the duty for which the machine is 
designed. 

“The efficiency is variable, and for the larger sizes is between 80 
and 86 per cent. falling off gradually as the pressures exceed 3 lb. 
per square inch. For smaller sizes the efficiency is less. 

‘‘Power for Rotary Blowers.—The horse-power required at the 
shaft or pulley to drive a rotary positive blower is proportional to 
the volume and pressure of the air discharged. It is safe to assume 
that for each 1,000 cu. ft. of free air discharged per minute at 1-lb. 
pressure, 5 h.p. is required. The following formule are sometimes 
used in calculating the horse-power. The first two formule give 
the theoretical horse-power required; and in order to determine the 
horse-power necessary to drive the rotary positive blower it is 
necessary to divide the results obtained by the efficiency of the ma- 
chine. The usual efficiency is between 80 and go per cent. 


er ee | 


11,000 


(Cr) a, 


“This formula is used when it may be assumed that the air is 
compressed so quickly that it does not have time to cool to atmos- 
pheric temperature, as in nearly all blower work. 


Oa = bs) 


33,000 


(2) hp. = 


52 AIR COMPRESSION AND TRANSMISSION 


“This formula is the ordinary “hydraulic formula” and is ordi- 
narily used for pressures up to 5 oz. 


lb. per Sa. ins x 
GQ) hpi v 


200 

“This formula is frequently used by makers of positive or rotary 
blowers for determining the horse-power required to operate the 
machine. In this formula Q represents the volume of air in cubic 
feet per minute displaced by the impellers, no allowance being made 
for slippage. In the above formule P, represents the pressure of 


= 

< 

‘5 0.06 
= Horse Power Curves. 

3 Machines of 1500 cv. ff. per 
ra min. capacity operating the 
20: samme tubes in the New York 
SS Postal Service, March 1909. 

+) 

c0) 

5. 

Ee 0. 

Le} 

oO 

AZ, 

ce) 

© 0. 


qui 


Horse Power Re 


afr Sees in ee Saves per square [heh 


Fic. 25.—Power consumed by rotary and piston compressor. 


the atmosphere or the suction pressure absolute in pounds per square 
foot and P the compression or discharge pressure in the same units. 
(See Fig..25.)” 


Mechanics of the Fan 


“The laws that govern the flow of gases are the same as those for 
the movement of liquids. If p, the pressure in ounces per square 
inch, is divided by 16 and this result multiplied by 144, the pres- 
sure will be expressed in pounds per square foot. This may also 
be done by multiplying d or the weight in pounds of one cubic 
foot of the gas by its height or head /# expressed in feet. 

That is: 


Ate = _ 9? 


and as the fundamental formula for velocity is v?= gh. 


v=V2 gh= 4/18 9 


PEGCOURPepyeV RSA DEI BYPARSONS 53 


“When # is given in inches of water: 


62.4 p=hd; na SP 


I2 


V=4/10.4 gt 


“The theoretical velocity obtained by using this last formula is 
greater than the actual velocity produced by the fan, because friction 
and eddies will restrict the freedom of flow. The formula, however, 
shows that the flow of gases through an orifice increases as the square 
root of the pressure and inversely as the square root of the density. 

“The head is made up of two parts—that necessary to overcome 
the friction and eddy losses and that necessary to produce the veloc- 
ity obtained. 

‘““The pressure produced by afan may be considered as equal to 
the weight of a column of gas one square foot in area which the fan is 
supporting. This weight is equal to the height of the column times 
the density of the gas. The “equivalent head’ is the height of 
this column of gas. Therefore, for any given pressure, the greater 
the head the less will be the density, and vice versa. Also, the 
greater the head required to produce a given pressure the greater 
will be the velocity. 

‘““As liquids have greater densities than gases, their equivalent 
heads for equal pressures will be less than the equivalent heads for 
gases. As velocities vary as the square roots of the head, the veloc- 
ity of gases will be greater than those of liquids under the same 
conditions of pressure. That is the reason why gases issue through 
orifices at greater velocity than liquids under the same pressure 
conditions. 

‘“‘As gases are compressible, their density will vary with the pres- 
sure. Their density also varies with the temperature and with the 
humidity contained. Since the velocity varies as the square root 
of the head, and as the head varies inversely as the density, any in- 
crease in density due to increase in pressure will reduce the head and 
consequently the velocity. 

“Conversely any increase in temperature reduces the density 
and consequently increases the head and also the velocity. The 
velocity is entirely dependent upon the head. Therefore, in mak- 
ing calculations for fan operations the effect of both temperature 
and density must be considered. For fan operation the standards 
generally adopted are: for temperature 60° F. and for density the 
weight of a cubic foot in pounds at atmospheric pressure or 14.7 lb. 


Therefore 


54 AIR COMPRESSION AND TRANSMISSION 


per square inch absolute. When the density of the gas is given 
for any pressure and temperature its density at any other pressure 
or temperature can be found with sufficient accuracy for all ordinary 
fan operations, by assuming that the density will vary inversely as 
the absolute temperatures, and directly as the absolute pressures. 


‘Thus d) = Z 
by 


“Tf a cubic foot of dry air weighs 0.077884 Ib. at 50° F. its weight 
uy (50-++460) X0.077884 


at 600° would be dy Ie es = 0.03751 lb. 
‘Also, as the density varies directly as the absolute pressure 
jab 
Pp 


“The pressure per square inch at atmospheric pressure is 14.7 lb. 
absolute or 235 0z. Therefore, the density at 3 oz. gauge pressure 
would be: 


_ (235+3) X0.077884 


dy 
235 


=0.0788 


“The head can be expressed for 50° F. when # is given in ounces 
per square inch thus: 


jp TAA ANP KP ROSS ee MPs 
6XdK 23ST IX(235+P) (235d-+dp) 
235 | 
‘“When # is expressed in inches of water as 
62.4 

vie 5-2 PK 400; a ae Ss eee 

ene PSEA) SE 

400.7 


“Since 62.4 lb. the weight of a cubic foot of water at 50° F. divided 
by 12 is the weight of a column of water 12 in. square and 1 in. high, 
and since the pressure at one atmosphere (14.7) would sustain a 
column of water 33.9 ft. or 406.7 in. high; 

“Substituting these values in the formula for velocity and ex- 
pressing p as the velocity pressure in ounces per square inch 


04 |26X ALLS Pi. HOA 32K 21 US pee [136036.8p 
235d+dp (235+p)d (235+p)d 


EECTURE BY MR. H. DE B. PARSONS 55 


When # is expressed in inches of water 


0=4)28X 2115p = (RE CEO | ee 
| 406.7d+dp N (406.7+p)d (406.7-+p)d 


For dry air at 50° F. d=0.077884. 


“These two formule are those used to calculate tables giving the 
theoretical velocities expected at different pressures. If the gas 
is not dry, but contains some vapor or moisture, its density will vary 


es Fahrenheit 


ie 
a 
Pe 
ee 
Fala 
a a 


Temperature of Air, Degre 


0 10 foe 30 40 50 60 70 80 
Vapor Content, Grains per Cu. Ft. (7000 Grains = | Pound) 


Fic. 26.—Humidity of air. 


by the quantity of mcisture which it contains, Fig. 26. (See Appen- 
dix C and chart.) 

“Tf the gas is at any other temperature than so” F. its density 
will decrease as the temperature rises, and conversely will increase 
as the temperature falls. As the temperature varies so will the veloc- 
ity. As the gas becomes lighter from increases of temperature, 
the velocity will increase as the square root of the ratio of the abso- 
lute temperature considered to the absolute temperature of 50° F. 
The converse is also true, as the gas becomes heavier from decreases 
of temperature, the velocity will decrease in the same ratio.” 


Effect of Outlet on Capacity 


‘The shape of the opening through which the gas is discharged 
from a fan affects the volume discharged in a given time. The 


56 AIR COMPRESSION AND TRANSMISSION 


shape of the orifice and the form of the duct affect the size of the vena 
contracta and therefore the blast area of the fan and the volume of 
the gas discharged. 

‘““As stated, the volume of discharge is a function of the product of 
the blast area times the velocity and the blast area is determined 
by multiplying the area of the orifice by the coefficient of efflux. 
The coefficients of efflux commonly used in practice for different 
types of orifices are: 


Orifice-in-a ‘thin plates... sn nraein ee a ee SSO 
Short cylindrcalspipessies cane eee eee ee eee ee O75 
Rounded ot-conicalmouth pieccss seat t eee 0.98 
Conical pipe, angle of convergence about 6 degrees..... 0.92 


‘With peripheral discharge fans, when the area of the outlet of a 
fan multiplied by the proper coefficient of discharge is less than the 
blast area of the fan, the pressure in the housing will equal that 
corresponding to the velocity of the tips of the blades, and the vol- 
ume of discharge will be less than the capacity of the fan. 

‘When the area of the outlet multiplied by its coefficient of dis- 
charge is greater than the blast area, the volume of discharge will 
be greater than the capacity of the fan, and the velocity of the gas 
as it enters the inlet must be greater than the speed of the inner edges 
of the blades. Consequently, the pressure in the housing will be 
less than that corresponding to the speed of the tips of the blades.” 


Work Required to Move a Volume of Gas 


‘A fan operating against 1 oz. of pressure per square inch and dis- 
charging the gas through an orifice having too sq. in. performs work 
which may be calculated in the following manner: 

“The total pressure against the fan is 100 sq. in. times the coeff- 
cient of efflux (say 0.75) times 1 0z., or 75 oz. or 4.7 lb. 

‘“‘ Assuming that the gas is air, then from the formula for velocity 
of dry air at 50° F. the air will have a theoretical discharge velocity 
through the blast area of 5,162 ft. per minute. 

‘The effective work is, therefore, 5,162 4.7 or 24,250 ft.-lb. 
per minute or 0.735 h.p. The actual work of driving the fan is 
greater than this result by the amount of power required to over- 
come the mechanical resistance and losses in the fan. This resist- 
ance is made up of the losses due to friction, windage and leakage. 
If these losses aggregate as much as the network then the power to 
drive the fan would be twice the network, 7.e., the efficiency of the 
fan as a machine would be 50 per cent. The actual power required 


to drive the fan would be Be 1.47 hip, 


EEUGLUKE BY VMK.Hs DEB, PARSONS 57 


“Placing the above in form of formule and taking f in ounces per 
square inch, 


Useful work =Py=—2 Hits 0S Fel See, 
Since jee 
16 


p _ dv 
De 18g 


From previous formula v?= 18 g 


adv 
Therefore useful work =—.— ft. lbs. per sec. 
2882 


‘““When # is given in inches of water 


Useful work =Py= 52” {t. lbs. per sec. 
144 
Since pete cee ued 
144°* 12 144 


p dv? 


From orevious formula v2 =10.4 ¢*, or p= 
: 4§ @ p 10.42 


5.2adv° ~—_—adv® 
I44X10.4g 288¢ 


Therefore useful work = ft. lbs. per sec. 

“In these formule, av is proportional to the volume of gas dis- 
charged by the fan. Since a is in square inches, the volume of cubic 
feet per second will be | 


av 
aaerr 


“Representing the efficiency by E the work to drive the fan may be 
fag eee es ah 
expressed as 238g Pols. Del sec, 

‘From these formule it will be seen that the power varies as the 
cube of the gas velocity, while the pressure varies as the square of 
the velocity and the volume directly as the velocity. 

‘From a consideration of these factors it is evident that fans are 
more economical when used to move large volumes of gas at low 
pressure than small volumes at high pressure. For this reason fans 
are not economical machines for compressing gases. In addition 
to the above, fans always have a clearance space between the 
revolving wheel and its housing, through which space the gases have 
a tendency when under pressure to leak backward, which tendency 
we have seen increases as the square root of the pressure. Fans are 
seldom used for pressures exceeding about ro oz., when higher pres- 
sures are desired the positive blowers are more efficient and are used 


58 AIR COMPRESSION AND TRANSMISSION 


for pressures as high as 8 lb. When still higher pressures are desired, 
compressors or blowing engines, such as described later should fa 
used. 

“Tn the formula above the value of v is that due to velocity head. 
When the dynamic head is known, that is the velocity head plus the 
friction or static head, a simple formula for brake horse-power to 
drive the fan is: 

‘““When ? is given in ounces per square inch 


ee 
brake h.p.= Bese 
“When # is given in inches of water 
OX5.2 p 
brake h.p.= ssoKE 


Design of Fans 


“Tt must be evident that unless a fan is properly designed for 
the work which it has to perform, there wil! be considerable loss in 
power required to drive it. 

“The peripheral speed of a fan must be such as to create the - 
desired pressure. The pressure against which the fan has to oper- 
ate is first determined, and having settled on the pressure the 
peripheral speed is made to conform with it. 

‘Furthermore, if the fan be direct connected either to an engine 
or to a motor, dhe speed of the fan will have to conform to that of 
the prime mover. 

“The work formule given above are all based on the blast area 
of the fan. The way in which these formule will apply to the differ- 
ent types of fans will be made clearer under ‘ Description of Fans.’ 

‘‘For any size of centrifugal fan there exists a certain maximum 
area over which a given pressure may be maintained, depending 
upon and proportional to the speed at which the fan is operated. 
If this area, sometimes called the ‘capacity area,’ ‘blast area’ or 
‘effective area of discharge’ be increased, the pressure is lower while’ 
_ the volume is increased. Contrariwise, if this area be decreased, 
the pressure remains constant while the volume is increased. In 
practice the outlet of a fan rarely exceeds the ‘blast area.’’’ 


- Description of Fans 


‘A disc, radial or propeller wheel fan consists of a machine having 
blades so mounted on an axle or shaft that when the shaft revolves 
these blades operate like a screw, and the gas is impelled ee 
in the direction of the axis. 

‘The blades may be straight and flat or curved. The blades may 
be curved in different ways so as to increase the screw effect and 


LECTURE BY MK. Ae DE B..PAKSONS 59 


diminish the centrifugal effect. Disc fans with curved blades will 
operate against slightly higher pressures and deliver more gas than 
those with straight blades. 

‘“‘As the gas enters the fan it will be forced forward with some 
centrifugal effect; and this centrifugal effect can be somewhat 
reduced by having the blades revolve inside of a tube so as to pre- 
vent the gas from escaping over the outer edges of the blades. 

‘“‘Disc fans will not operate economically against a pressure, as 
the pressure will increase the slip and the leakage of the air from the 
blades at the tip. If the pressure is at all high the gas will be drawn 
backward near the axis and will be blown forward near the outer tips 
of the blades, or in other words, the fan disc will simply circulate 
the air without making any delivery. 

‘Disc fans operate best when drawing gas from a practically free 
suction and discharging it at no pressure. When these fans are 
set up care must be taken that they do not operate against the wind, 
as the wind pressure will vitiate the operation of delivery. 

“The number of blades appear to have a small effect upon the 
discharge, provided, of course, that the number is neither too large 
nor too small. Too many blades will simply churn the air and 
produce the effect of cavitation. Too few blades will not give a 
sufficient grip on the air to force it forward with the proper 
delivery. 

‘The gas delivered by a disc fan is very irregular in velocity. If 
anemometer readings are taken at different points in front of the 
disc, the recorded velocities will be found to vary without apparent 
reason, and the variation will not remain constant. It is, therefore, 
very hard to determine the mean velocity of discharge of the gas. 

“The number of revolutions is limited by the strength of the fan 
and by the fact that a high velocity will cause the fan to hum and 
be noisy. The revolutions are usually limited so that the velocity 
at the tips of the blades shall not exceed 8,500 ft. per minute. (For 
noiseless operation, 4,000. Usual maximum 7,000). On this as- 
sumption, if D denotes the diameter of a fan in inches and 


nm denotes the number of revolutions per minute. 


zDn 
es see sures and Dn=32,000 (nearly). 
“This last equation may be used to determine the limiting revo- 
lutions or diameter by assuming one or the other. 
“The volume of gases discharged by a disc fan with a free suction 
and discharge can be estimated from the formula: 


D? : 
=1/4 2—v in which v=0.20V 
Q=1/4 a 39 


60 AIR COMPRESSION AND TRANSMISSION 
“The brake horse-power for the fan with the above value of Q 
can be estimated from the formula: 


_OXxdXv3/? ( 13.5 is a constant 
RLS OED 550X 28 ‘S \found from experience 


‘When a disc fan operates against a pressure, 7.e., not with a free 
suction and discharge the above formule must be changed as Q 
becomes less because the slip becomes greater. 

2 

‘‘ Approximately 0 = 1/47 (=) vy in which v! equals 1.25v less 45 
per cent. of the theoretical velocity due to the pressure against 
which the fan is working. 

“The brake horse-power will be the same as if the fan were work- 
ing without restriction, although the volume of discharge Q will be 
less. 

“Example: Free suction and discharge. Fan wheel 48 in. in 
diameter running at 450 revolutions per minute. Find Q and 
brake horse-power. Dry air at 50° F. 


a Se ee ee 
V=0.39X3.14X— Xoo 36.8 


8 2, 
Q=1/4X3.14Xx (E) X36.8=462 


462 X0.077884 X 36.83/2 X 13-5 
550 X 64.32 


Brake horse-power = 3:07 

‘‘Also restricted discharge. Find Q and brake horse-power for 
the same fan and conditions when operating at 5/8 in. of water 
pressure. 

“The velocity due to 5/8-in. water pressure is 51.8 ft. per second. 


vi=1,.25 X36.8—0.45 X51.8= 22.8 
Q\ 2 
Q=1/4X3.14X (~) X22 9 = 250 


“The brake horse-power would be 3.07 because it is approximately 
the same as if the fan were working unrestricted. It is found by 
substituting the unrestricted value of Q instead of the actual re- 
stricted value. 

“Centrifugal Fans.—Centrifugal fans operate on the principle of 
the vortex. They suck the gas in and discharge it off the periphery 
of the wheel by centrifugal action. 

‘Fan Blast or Steel Plate Machine.—The fan wheel consists of 
an axle or shaft on which are mounted radial arms carrying floats or 
blades. Each blade is narrower across the tip than it is across the 


LECTURE BY MR. H. DE B. PARSONS 61 


body. The blades are mounted inside of side plates, so that the 
gas is confined in the spaces between the blades, which thus form 
passages from the suction to the discharge side of the fan. Thése 
side plates also prevent the loss of friction between the revolving 
air and the sides of the housing. 

‘“‘Sometimes the blades are curved backward at the tips so as to 
make the fan run more quietly, and sometimes the blades are curved 
backward for their whole depth so 
that the gas may enter the wheel 
and pass through it without shock. 

‘““When fan wheels have flat 
blades, they can be run equally 
well in either direction, but when 
the blades are curved, the wheels 
should revolve with the convex 
sides of the blades in advance. 

‘““When these fans are used for 
blowers, there is usually an inlet 
on both sides of the housing; and 
when used as exhausters, usually 
an inlet on one side only, as it 
facilitates the connection with the 
suction duct. w+ 04D 


es 


4 

‘ 

‘ 

J 

“The diameter of inlet is gener- wid ae 9 

_ 1 
ally 0.6 or 0.7 of the diameter of 4-4 /east ia, but beection 
Ahad heal actually from ' of Fan 
e Ian wheel. 2a to $a. : Wheel 

“For high efficiency the area of Lv 


inlet should not exceed 4o per cent. 
Ol athe dise, areayo! they wheel. 
The full width of the blades is 
generally made either one-half or three-eights of the diameter of 
the “wheel. The blades are generally cut off at the upper outer 
corners so as to taper at the tips at an angle of about 20 degrees 
with the side edges, but their width at the periphery should be not 
less than 0.6 to 0.7 of the width at the root, 7.e., not less than their 
maximum width times the same ratio as was chosen with the ratio of 
inlet to wheel diameter. Usually w=o0.4 D and w=o.5 D (see 
Fig. 27). | 

‘The width of the fan is made such as to provide the proper area 
for the flow of the gases through it so as to discharge the required 
volume. If the diameter of the fan wheel is made too small, it may 
not be possible to give the wheel sufficient width to permit the 
necessary discharge of volume, unless the fan is run at a very high 
rate of speed. This increased speed will result in raising the pres- 
sure above that required, and will, therefore, increase the power 


Fic. 27.—Steel plate fans. 


62 AIR COMPRESSION AND TRANSMISSION 


necessary to drive the fan. Contrariwise, if the wheel be given 
a large diameter it may have to be made so narrow, in order to dis- 
charge the required volume, as to become impracticable. It will, 
therefore, be seen that under any given conditions there will prob- 
ably be one diameter and width which will be best suited for the 
work. 

“The blades of this type of fan are given sufficient depth so as to 
project inside of the circle of the inlet in order that they may better 
grip the incoming gas and force it through the wheel. 

‘With a peripheral discharge fan enclosed in a housing, the limit 
of its capacity to maintain a given pressure is measured by its blast 
area. In other words the velocity of discharge will be approxi- 
mately equal to the peripheral speed of the fan, and the volume will 
be measured by this velocity times the blast area. 

‘“‘Tf the blast area be increased the pressure will be less, and if the 
blast area be decreased the pressure will remain the same. For a 
peripheral discharge fan with a housing the blast area can be cal- 
culated as follows: 

Let D denote diameter of fan wheel in inches; 

w denote width of fan wheel at periphery in inches; 
c denote a constant, depending upon the design of the fan 
and its housing, but which has a value not far from 2 1/2 to 


? 
a denote the blast area in square inches. 


D D 
Blastrarea—o — es = = g(nearly) 


“Tf the shape of the discharge orifice and duct be known, and the 
coefficient of contraction determined, the area of the discharge orifice 
would be the blast area, as determined from the above formula 
multiplied by the reciprocal of the coefficient of contraction. 

“The usual maximum peripheral velocity for standard fans is 
6,600 (for noiseless operation about 4,200) ft. per minute, but should 
not exceed 8,000 ft. per minute. This latter figure limits the pres- 
sure to 1 3/4 oz. per square inch but special fans may be designed to 
maintain a pressure as high as about 12 oz. 

“The volume past the blast area is about 86 per cent. of the 
peripheral speed. In other words, the peripheral speed must be 
1.16 times the velocity due to the pressure of V =1.16 2. 


V 
wheel circumference 


Therefore n 


“The efficiencies without prime movers vary from 45 to 50 per 
cent. for commercial sizes when using dynamic head, or from 30 to 35 


EPEGCLOREOBY MRA Ho DE B- PARSONS 63 


per cent. when using velocity lead. The outlet is generally made 
square, and its area is usually about two and one-half times the 
blast area, or A =2 1/2 a, but never less than 1 1/2 a. - 

“This proportioning will make the bottom of the outlet below the - 
periphery of the wheel. The efficiency of commercial sizes is about 
45 to 50 per cent. without a prime mover. If the prime mover 
efficiency is taken at 85 to 90 per cent. then the total efficiency of 
the fan and prime mover would be between 38 and 45 per cent. 
Example: 

‘“‘Given the quantity of air per minute, 65,000 cu. ft., the temper- 
ature of dry air 70° F. and the pressure 1 3/4 in. of water. Deter- 
mine the diameter of fan, revolutions and brake horse-power. 

Under these conditions the density of the gas or its weight per 
cubic foot may be taken as 0.0754 lbs. 


a] 136036.8 X1.75 e 
v= = 88.0 
(406.7-+1.75) X0.0754 
65000 «X88.0 _ 65000X144 _ 
eee: Therefore a= 60% 88.0 =1,770 


Making w=0.4D 


dak 
1 7O= 91 D*=11,500,, D=107 


ez", 10 6 00.0 == 102 


Wheel circumference x14 = 27.0 it. 
ESE 
27-9 
With E as 45 per cent. the power to drive the fan is 


Therefore nN 220 


1770 X0.0754 X 88 By 
550X288 X 32.16 X0.45 


“‘Ffousing.—The housing is placed around the wheel in an eccen- 
tric position and has a form approaching the spiral. This arrange- 
ment facilitates the gas delivery from the wheel. The openings 
for discharge of the gas are tangential to the wheel. There may be 
one or more openings as circumstances demand but their combined 
area of discharge should not exceed the fan capacity. It makes no 
difference whether these discharge outlets are placed horizontally 
or vertically. 

“The arrangement of discharge outlet, however, gives a name to 
the fan—as a horizontal top discharge, a vertical discharge, a hori- 


Brake horse-power 


64 AIR COMPRESSION AND TRANSMISSION 


SS 
SS 
S wf 


——— 
SS Ss 
SSS Ss 
=" 
= 


SSeS 
See 
Se 
SSeS sss 


Ss 


Sse 
SS 


== 35553 [22SS 
= = 

SSS 
ee 


Fic. 28.—Full housed steel plate Fic. 29.—Full housed steel plate fan. 
fan. Left-hand bottom horizontal Right-hand top horizontal discharge. 
discharge. 


Fic. 30.—Three-quarter housed steel Fic. 31.—Three-quarter housed 
plate fan. Right-hand bottom horizon- steel plate fan. Left-hand top 
tal discharge. horizontal discharge. 


LECTURE BY MR. H. DE B. PARSONS 65 


zontal bottom discharge, a double discharge, etc. (Figs. 28, 29, 30, 
2randes 2.) 

“The spiral or scroll form of the casing should be such as to let 
the gas escape with freedom from all parts of the periphery. The 


smaller diameter of the scroll should not be less than Di and the 


re ence 


D denotes the diameter of the wheel in inches; 
a denotes the blast area in square inches; 
W denotes the maximum width of blades in inches. 


larger diameter not less than D+ 


de 
fae 


O oy” 
A! 


qu 
Fic. 32.—Allis Chalmers steel ventilating fan. 


‘Cone Wheel Fans.—The cone wheel fan is a single inlet pe- 
ripheral discharge fan. It is used both with and without a housing. 
Cone wheel fans are not efficient for use against pressures in excess 
of 1 oz. per square inch and are seldom used against pressures as high 
as this limit. Generally speaking, they are not as economical in 
the handling of gases as centrifugal fan-blast machinery properly 
encased in a well-designed close-fitting housing. 

‘““Cone wheels should have a perfectly free inlet and be arranged 
to have a free discharge of air from all points of the periphery. 
When cone wheels are encased in a housing the housing is usually 
much larger than the fan wheel to permit a perfectly free and un- 
restricted discharge. As ordinarily arranged, the inlet to a cone 
wheel is a hole in a wall of the apartment from which the gas is to be 
sucked-(Figs. 33 and 34). 

“On the axle or shaft of the fan there is mounted a cone with its 
apex turned toward the inlet. Between the cone and the periphery 
of the wheel there are blades or floats, and these blades are encased 

5 


66 AIR COMPRESSION AND TRANSMISSION 


inside of side plates. As the air enters the inlet it is deflected by 
the cone to the floats, which together with the side plates, continue 
to change the direction of the air so that it is discharged off the pe- 
riphery in a plane at right angles to the shaft or line of entrance. 
The width of cone wheels is generally one-quarter the diameter of 
the wheel, and the inlet opening is generally three-quarters of the 
diameter of the wheel. The floats are curved backward and tapered 


WE 


<a 


Yt] 


Fic. 33.—Cone fan inlet side. Fic. 34.—Cone fan discharge side. 


toward the periphery so that they have a width at the tips of about 
three-quarters the width of the wheel. 

“Assuming that the air is discharged at a velocity equal to the 
speed of the tips of the floats, the capacity of a properly designed 
cone wheel in cubic feet per second is approximately: 


QO=6.4D%*/p 
for exhausting and O=s.0D?\/p for blowing in which 


D denotes the diameter of the wheel in inches and 
p denotes the pressure in ounces per square inch corre- 


sponding to the velocity of the tips of the blades. 
‘The horse-power required to operate a cone wheel including 


efficiency is approximately brake horse-power = for exhausting 
43 


and 


brake horse-power=? for blowing. 


PECTOREPUBY MK ae DE BeePARSONS 67 


“The limits of peripheral speed are about the same as for a disc 
fan so that assuming the speed of the tips of the blades, the revo- 
lutions and diameter can be calculated by assuming one or the other 
as with disc fans. 

“Turbine Blast or ‘Sirocco’ Fan.—The name Sirocco is a trade 
name. ‘These fans are centriufgal in their action and have a pe- 
ripheral discharge. ‘The runner or blast wheel is built up of steel, 
and consists essentially of three parts—the interior cone to deflect 
and turn the air as it enters the inlet toward the blades, the blades, 
and the side plates. The runner is shaped like a drum. The 
blades are Jong and narrow radially, being generally between six 
and nine times as long as they are wide. The runners are usually 
equipped with about 64 blades. The blades are curved so that the 
concave side revolves forward. The blades are not quite as deep 
as the side plates, and the side plates are made one-sixteenth the 
diameter of the wheel. 

‘““These fans are usualy made with the inlet on one side only, the 
other side being closed by the back of the internal cone. When 
double inlets are desired, two internal cones are placed back to 
back, and this practically means that the wheel is made double. 

‘‘The peripheral speed at which these fans can be run before they 
begin to hum or become noisy is higher than that of the steel plate 
fan. With ordinary sizes they remain quiet until peripheral speeds 
of 10,000 ft. per minute are reached, but a peripheral speed of over 
15,000 ft. per minute is often used. 

“The design of these fans is quite different from that of a steel 
plate fan. Having the blades curved forward, which is quite correct 
in principle, results in an increase in the velocity of discharge, which 
with properly shaped blades, is twice that of the “‘steel plate’’ fan. 
In other words, the velocity of discharge from a properly designed 
turbine fan, is theoretically twice that of the periperal speed, and 
actually about 117 per cent. of the peripheral speed. As these fans 
are made about twice as wide as the steel plate fan, and as the 
velocity of discharge is twice as great, a turbine fan wheel will dis- 
charge about four times the volume of gas as a steel plate fan of 
equal diameter and run at equal speed. 

‘These fans are mounted in a housing, and the ratio of outlet to 
inlet is one totwo. There are three types of these fan wheels which 
are distinguished by the manner in which they are bladed (Fig. 35). 

(a) In the forward inclined wheel the blades are curved 
forward of the radii. 

(b) In the radially inclined wheel, the blades are curved on 
the radii. 

(c)’ In the backward inclined wheel, the blades are curved 
behind the radii. 


68 AIR COMPRESSION AND TRANSMISSION 


“These different forms rank in the order of mechanical efficiency 
as given above, and in the matter of speed of revolutions in the 
reverse order. Owing to the high speed at which these turbine fans 
can be run they are well adapted for direct connection to high-speed 
prime movers (Fig. 36). 

“The commercial efficiencies when used for blowing are about 59 
per cent. for forward blades, 53 per cent. for radial blades and 50 
per cent. for backward blades. 


SSS 
WNZE NZ 


C 
Fic. 35.—Three types of blading. Fic. 36.—Sirocco double inlet runner. 


‘The width of the runner is made two-thirds the diameter for 
standard sizes. 
‘The volume of discharge can be calculated from the formula 


av 
=———. in which y=V X11; 
Car v, 44 
2) Dae met 
a SS Se eee en 
277 aes] SO 225 
‘‘When the dynamic head is known the same formula for power 
required can be applied to this fan as to others.” 


CHAPTER VII 


PISTON COMPRESSORS 


In the practical applications of air for power and other purposes, 
the simplest method of compressing air above those pressures 
for which the fan and positive pressure blower are particularly 
adapted is by means of a piston compressor. 

This method of compression is used more than any other method 
and for that reason considerable attention will be given to it. Fig. 
37 shows in cross-section such a compressor, its piston, piston-rod 


\ IZA i 
A 
ER SC 
Kea LS 


Of LLIN 


Soom 


iS 


lf 


LYtEY 


t 


le 
aes) 


RNewa 
[DOISSSSSSS 


| Ze 


iY, 
Sinn 


Fic. 37.—Cross section of piston compressor. 


and valves. D represents the inlet valve through which the “free 
air’? is drawn into the compressor, and E the outlet or discharge 
valves through which the compressed air is discharged into a reservoir 
or receiver as it is quite commonly called. 

Naturally with this type of compressor the piston movement 
is limited, so that at its extreme positions the piston will not strike 
the cylinder ends or heads. That volume of the cylinder through 
which the piston does not move, together with the volume occupied 

69 


70 AIR COMPRESSION AND TRANSMISSION 


by the passages leading from the cylinder to the valves, is called the 
clearance. 

Action of Piston Compressor.—Suppose the piston is to start 
at the right end of its stroke. As it moves to the left, a slight 
vacuum will be created in the cylinder on the right side of the piston 
and valve D will be opened. Free air will rush in, following the 
movement of tke piston and filling the cylinder. As the piston 
starts to the right, this same operation will take place on the left 
side of the piston, while on the right compression will commence as 
the volume occupied by the air in that part of the cylinder is reduced. 

If the valve E is in communication with a reservoir of com- 
pressed air at 30-lb. gage pressure, no air can escape from the cylinder 
until the pressure has risen to a little above 30 Ib. When this is 
done the valve E will be lifted from its seat and the compressed air 
will be pushed bodily out of the cylinder into the reservoir. 

At the end of the stroke, the clearance volume will be filled with 
this compressed air and as the piston starts back valve EF will close 
and the compressed air in the clearance space will expand to fill 
the gradually increasing volume in the cylinder and will continue 
to do so until the pressure in the cylinder is lower than that of the 
atmosphere, when valve D will open and a new supply of free air 
be drawn in. The clearance space reduces the volume of free air 
‘ drawn into the compressor. 

Indicator Card of Piston Compressor.—These changes are shown - 
by the indicator card given in Fig. 38, which shows the changes in 


| 
| 
| 
| 
| 
| 
! 
! 
I 
j 
r 
| 
| 
| 


[aie 


cap) 


V 0. 


Fic. 38.—Indicator card of piston compressor. 


pressure and volume taking place on one side of the piston. The 
distance LZ represents the volume displaced by the piston during 
one stroke. Starting with the piston at G and the cylinder full of 
air, the pressure of the confined air will gradually rise as the volume 
is being reduced by the moving piston until a pressure H, a little 
above the pressure in the reservoir is reached. From here until 


PISTON COMPRESSORS 71 


the end of the stroke the piston will force the compressed air out 
of the cylinder into the reservoir. The wavy line H—J represents 
this expulsion, the inertia of the moving parts of the indicator and 
the fluttering of the discharge valve on its seat causing the irregu- 
larities of this discharge line. When the stroke is completed the 
pressure of the air in the reservoir closes the valve F, leaving the 
clearance space full of compressed air at the high pressure J. As 
the piston moves to the left the compressed air in this clearance 
space will expand to fill the constantly increasing volume until a 
pressure a little below that of the atmosphere is reached (K), when 
valve D will open and free air rush in to fill the cylinder as the piston 
continues its movement. 

The more air in the clearance space the further wil] the piston B 
have to move before the compressed air in the clearance space can 
have the opportunity to expand low enough to open the inlet valve. 
For this reason the clearance space for a piston alr compressor 
is made as small as possible. 

Effect of Clearance.—The loss due to clearance is not a loss of 
power, for most of the energy used in compressing the air into the 
clearance space is given back in expanding and helping to move ‘ 
the piston. The only loss in power is the heat loss through radia- 
tion. The loss due to clearance is mainly a loss of capacity which, 
in many cases, is a rather serious matter. Engineers have sought 
to reduce this to a minimum with the result 
that the clearance volume of a modern piston 
. air compressor varies from 0.02 to 0.0094 of 
the volume of the piston displacement. 

Methods of Reducing Clearance.—Various 
methods have been devised to reduce this 
‘ clearance, some even going to the extent of SEES WY 
putting in spring heads on the cylinder which MMM LL 
the piston could strike at each end of the F'6. 39.—Uncovering 
stroke without serious injury. This method, ? PEA EARL gee 

: . Bing pressure. 
however, introduces complications that are 
not always desirable. 

Another method of reducing the clearance loss that has been 
suggested is to let the piston uncover a passage Jeading from the 
clearance space just at the end of the stroke, and thus allow the 
compressed air in the clearance space to expand into the cylinder 
on the other sid: of the piston, as shown by Fig. 39, without reducing 
the capacity of the compressor. This, however, is open to the 


72 AIR COMPRESSION AND TRANSMISSION 


objection that the compressed air in the clearance space, which 
normally acts as a spring, is released and the piston will pound 
at the end of each stroke unless some other means is used to prevent 
it. 

Some piston air compressors are so designed that when the 
machine is cold the piston will almost touch the cylinder-head 
when at the crank end of its stroke. As the compressor is operated 
the heat of the air being compressed will be partly transmitted to 
the piston-rod and cause it to expand slightly, thus increasing 


Cc B B 
LZ, ae. IZ, eae p, K 

== |") 
Z % 

y y 

4__y 

= 

ZZZZLLLETLE y 

ee 


sxx 


oe 


Fic. 40.—Hydraulic piston compressor. 


the crank-end clearance and decreasing the head-end clearance, 
and by designing the compressor so that when hot, the piston will just 
miss touching the cylinder-head when at head-end dead-center, 
and when cold just miss the other cylinder-head when at dead- 
center, the clearance volume of the cylinder is reduced to a minimum. 

Some of the early air compressors were built, as indicated by 
Fig. 40, so as to reduce the clearance volume by using a water column 
as a piston. As the piston A is moved back and forth the water 
column in each upright cylinder is caused to alternately rise and 
fall. As it falls, valve C is opened and free air rushed in. As the 
water column is raised, valve C is closed and the confined air is 
compressed to a pressure sufficient to open valve B and permit 
the compressed air to escape to a reservoir. The water at its upper 
height can fill the entire space, including the passage to the valves, 
which a metal piston could not do, so the clearance is in this way 
reduced to a minimum. This type of compressor was practically 


PISTON COMPRESSORS 73 


abandoned, but its principle has been recently revived in air com- 
pressors working on the principle of the Humphrey pump. 

Suction Line.—If in the design of a piston compressor the inlet 
valve or the passages for the same should be too small, then the air 
cannot rush in as fast as the piston 
moves and the suction line, instead 
of being straight, will fall below the 


horizontal, to rise again near the StS. 
end of the stroke as the piston 
- ee 
velocity decreases. Ne 


A suction line that is not hori- 
zontal indicates restricted inlet for 
admission of air. 

It may sometimes happen that as the piston, which has its max- 
imum velocity near the middle of the stroke, nears the end with a 
decreasing speed, the inlet valve will close before the end of the 
stroke is reached, and the admission line will fall slightly. As 
compression starts, this line will be retraced until a pressure greater 
than the admission pressure is reached, as shown by Fig. 41. A 
good indicator card for a piston air compressor will have the com- 
pression line start very close to the end, as shown by Fig. 38. 

Compression Line.—One reason for the ideal method of air 
compression being isothermal, as already explained, is because 


Fic. 41.—Effect of early closing of 
inlet valve. 


Pp 


0 V 


Fic. 42.—Card showing isothermal and adiabatic compression. 


any energy stored as heat in the compressed air above the temperature 
of the surrounding atmosphere will soon be radiated and hence lost. 
Another reason, as shown by the chart, Fig. 11, is that if the com- 
pression is not isothermal, the pressure due to the increased tempera- 


74 AIR COMPRESSION AND TRANSMISSION 


ture will rise above that which would result from isothermal com- 
pression and hence cause an increased expenditure of energy to 
operate the compressor, as.shown in Fig. 42. This increased 
expenditure of power may be avoided by isothermal compression 

The approximate horse-power required to compress air under 
isothermal or adiabatic conditions may be determined as indicated 
by the formule in Chapter IV. This has been done for various 
pressures indicated in Table IX, which will give an idea of the 
saving to be secured from isothermal compression. 


TABLE IX 
Single-stage compression, from atmospheric pressure at sea- 
level. Inital temperature, 60° F. Horse-power required 
to compress 1 cu. ft. of free air 
Atmos- 
Gage | pheres Calculated horse-power Actual horse-power (approx.) 
pres- | absolute . 
qhien aCe Sac Allowance for | Allowance for 
pounds} of com- 
: : : losses above losses above 
pression | Isothermal | Adiabatic ; : j ; 
; ‘ adiabatic com- | adiabatic com- 
compression | compression 3 : 
pression, 15 per | pression, 20 
cent. percent: 
20 2.30 0.0551 0.0626 0.0720 0.0751 
25 AR ie 0.0637 0.0741 0.0852 0.0890 
30 3.04 ORO7 Ts 0.0843 0.0970 ©. LODI 
35 3.38 0.0782 0.0941 0.1082 ©.II29 
40 aa72 0.0842 0.1029 oO, T1S2 0.1234 
45 4.06 0.0895, Ovrrir 0.1282 0.1338 
50 4.40 0.0950 OnlLor 0.1470 0.1430 
55 Ant A 0.0994 0.1269 0.1460 Oris22 
60 5.08 O.IO4I Or1337 0.1537 oO. 1604 
65 5.42 0.1081 ©. 1401 0.1610 0.1681 
70 5.70 Omer res 0.1468 ©. 1690 0.1761 
75 6.10 O.1162 O7E535 0.1765 0.1842 
80 6.44 O.IIQ5 On1s0n 0.1830 ©.I9I0 
85 6.78 Ov1224 OBTOSE 0.1900 oO. 1961 
go ipl Ont250 CO: 1702 0.1955 ©. 2040 
95 7.46 0.1287 0.1760 0.2024 O;2112 
100 7.80 eyes 8 0.1807 0.2080 0°. 2168 
IIo 8.48 0.1366 0.1894 0.2180 O, 2272 
125 9.50 ©.1442 0.2025 0.2328 ©. 2430 


Wet and Dry Compression.—The two principal systems which 
have been used in attempting to secure isothermal compression in a 


PISTON COMPRESSORS | 75 


piston air compressor are the ‘‘wet system” and the “dry system.” 
A. wet compressor is one which introduces water directly into the 
cylinder during compression. 

A dry compressor is one which admits no water to the air during 
compression but surrounds the cylinder with a jacket of circulating 
water in order to reduce the heat of compression. 

There are two kinds of wet compressors: First, those which in- 
ject water in the form of a spray into the cylinder during compres- 
sion; second, those which use a water piston in compressing the 
air, such as shown in Fig. 4o. 

Numerous tests have been made of these different methcds of air 
compression showing that the compression cen be brought closest 
to the ideal isothermal by means 0i injecting a spray of water directly 
into the cylinder during the compression. 

Although the best results have been secured by the wet system of 
compression, still it has been practically abandoned in favor of a 
dry system of compression using a water-jacket, for the following 
reasons: 

First. The mechanical difficulty of introducing the water in a 
fine enough spray to reduce the temperature of compression as It is 
being produced. 

Second. Impurities in the water through mechanical and chem- 
ical action destroy the metallic surfaces of the cylinder and piston. 

Third. Wear due to insufficient lubiication. 

Fourth. Difficulty of regulating the amount of water to be 
introduced. 

Fifth. Limitations of speed due to the presence of water. 

Actual Compression.—If an isothermal and an adiabatic line be 
drawn on an indicator card taken from an actual modern air com- 
pressor starting at the point where compression begins, the actual 
compression line wiJl come very close to the adiabatic. 

If the compressor operates at a very slow speed, there is an 
opportunity for the heat that is generated by compression to be 
radiated and the compression line will come closer to the ideal 
isothermal. 

That is, with a high-speed air compressor, the compression is 
approximately adiabatic, while with a slow-speed compressor with 
efficient water-jacket the compression line may approach the 
isothermal, speed being an important element in determining the 
slope of the compression line. 

Cards from Air Compressors.—In taking indicator cards from an 


76 AIR COMPRESSION AND TRANSMISSION 


air compressor all the precautions that are necessary for taking 
steam cards apply with equal force, as erroneous conclusions may 
be very easily drawn from the cards if care is not taken. For 
instance, if the piston of the air compressor leaks slightly, then the 
compression line will approach the ideal isothermal line, indicating 
' a very desirable compression line, while in reality the compressor is 
defective. For this reason, extra precautions must be taken in 
order that the card may indicate the true state of affairs in the 
cylinder. 

When air is compressed in a cylinder to a pressure of 100 lb. per 
square inch without cooling, temperatures ranging from 475° to 
550° are reached, and these high temperatures are not only produc- 
tive of poor economy and low efficiency but are dangerous because 
of the explosive nature of the compressed air containing the vapors 
of the cylinder oil used for lubricating the piston. 


CHAPTER VIII 


EFFICIENCIES AND ENERGY COMPENSATION 


In the discussion of Chapter IV, clearance was disregarded, but in 
calculating the required dimensions of an air compressor the effect 
of clearance must be considered. 

It is usual to express clearance as a certain percentage of the piston 
displacement. If this percentage expressed as a decimal fraction 
is represented by C, the volume occupied by the air to be com- 
pressed at the end of a suction stroke will be (1+C) times the 
piston displacement. — 

Volumetric Efficiency.—The effect of clearance upon the capacity 
of a compressor js usually expressed in terms of the “volumetric 


C B 


' 
I 
' 
{ 
|! 
1 
i 
! 
! 
1 
i 
1 
| 
1 
' 
G 


0 


Fic. 43.—Ideal card for piston compressor. 


efficiency,” but as this term is not always interpreted in the same 
way it is advisable to use two terms “apparent volumetric efficiency” 
and “‘real volumetric efficiency.”’ 

Apparent Volumetric Efficiency.—The apparent volumetric eff- 
ciency is the apparent volume of free air drawn in as shown by 
’ the indicator card divided by the volume of the piston displacement, 


Oralieice. a AG of ideal card, Fig. 43. In an actual card, Fig. 44, this 


AD 
ratio is also shown by 1G" 
Hi 


78 AIR COMPRESSION AND TRANSMISSION 


In Fig.-43, the clearance line C—D will follow the equation 
beVe" = paVa” 
but, as V-=C, this may be written 


pPeC” =paVa"™ 


be\n 
va=(F) C 


from which 


Fic. 44.—Actual card of piston compressor. 


The apparent volumetric efficiency, or en may be written 


AG—DG 

AGES 

or id 
DG 
DEINE 

or 

pee 

AG 


Calling the piston displacement AG unity, the apparent volumet- 
ric efficiency may be written 


I 


Be ; Pe \n 
_-ofts)ixccex-ef (22) 


The effect upon the capacity may be illustrated by assuming a 
compressor in which #¢ is 80 lb. per square inch gage, or 94.7 lb. ab- 
solute, and C is 2 per cent and 7 is 1.4. 


EFFICIENCIES AND ENERGY COMPENSATION ico 


Substituting in the above, 


| for Three Stage 


pe\eu 
Po 


| for Two Stage ( 


14 
“| for Single Stage 


oh 


0.5 


Fic. 45.—Loss of capacity due to clearance. 


That is, such a compressor would only take in 94 per cent. of the 
piston displacement in free air, and poor valve action would reduce 
this capacity still further. The loss of capacity due to clearance for 
various pressure ratios is shown in Fig. 45. 


80 AIR COMPRESSION AND TRANSMISSION 


True Volumetric Efficiency.—The above illustration would be 
true if the temperature of the air after being drawn into the cylinder 
were the same as the atmosphere, and the pressure at the instant of 
compression equal to the atmosphere. As this is seldom true, it is 
necessary to make correction for this by multiplying the above 


expression by 
ae A 
iP Pam 
in which the subscript am stands for atmospheric conditions and 
subscript 1 for conditions at the beginning of compression. 
The true volumetric efficiency is the ratio of the free air actu- 
ally drawn in to the piston displacement and is represented by the 


formula 
Alas Pi Pe ae 
Tie lad(enie |! 


Cylinder Efficiency.—The cylinder efficiency of an air compressor 
may be defined as the ratio of the work done in a complete cycle to 


> 


Fic. 46.—Cylinder efficiency. 


compress isothermally a volume of air at atmospheric pressure equal 
to the intake piston displacement divided by the actual work done 
in the air cylinder. 

This would be Fig. 46, the area AKCG divided by the shaded area, 
or the actual work done in the air cylinder. 

Efficiency of Compression.—The efficiency of compression may be 
defined as the product of the cylinder efficiency and the true volumet- 
ric efficiency, or it is the work done in a complete cycle to compress 
isothermally, without clearance, a given volume of free air divided 
by the work actually expended in compressing the same volume of 
free air. 


EFFICIENCIES AND ENERGY COMPENSATION 81 


Mechanical Efficiency.—The mechanical efficiency of an air com- 
pressor is the work done in the air cylinders divided by the work done 
in the steam cylinders, if driven direct by steam, or in the gas-engine 
cylinders, if gas engines are used, or the work delivered at the belt 
if the compressor is belt driven. 

Net Efficiency.—The net efficiency of a compressor unit driven by 
a steam engine or turbine direct is the ratio of the internal energy 
available in the compressed air at room temperature to the heat 
energy available in the steam supplied; or it is the energy available by 
adiabatic expansion of the compressed air at room temperature to 
atmospheric pressure divided by the energy available in the steam 
supplied, if expanded adiabatically in a Rankine cycle. 

In considering efficiencies of air compressors, it is important to 
distinguish between a machine used for compressing air as a means 
of storing and transmitting mechanical energy, in which the ideal 
compression is isothermal, and a machine used for supplying air 
under pressure for purposes of combustion, as in forges, cupolas and 
blast furnaces. In these last cases the pressures are comparatively 
low and the resulting increase of temperature due to adiabatic com- 
pression is not objectionable. In fact there is ample justification 
for taking, in these cases, adiabatic compression as the standard. 

Blower Efficiency.—Henry F. Schmidt in an article in the Journal 
A. S. M. E. of Nov., 1912, on “‘Centrifugal Blowers” indicates a 
“blower efficiency’? for any blower not water-jacketed, by dividing 
the rise of temperature, as calculated from adiabatic compression 
from the suction to the discharge pressure, by the actual rise of 
temperature taking place during the compression in the blower. 

The losses in a blower are principally friction, eddies and leakage. 
All energy losses reappear as heat and bring the temperature after 
compression higher than that due to adiabatic compression, and in 


the article the author proves that this ratio will reduce to the form 
11—T2. ; ae ; 
ToT, in which Tis the initial temperature of the air, 71’ its actual 
final temperature, and 7, the final temperature if the compression 
had been adiabatic. This formula is open to the criticism that the 
radiation is disregarded, but as its value is comparatively small the 
‘blower efficiency” expression has the decided advantage of sim- 
plicity and ease of determination. 

Economic Efficiency.—Franz zur Nedden in his articles on Turbo- 
blowers and Compressors in the Engineering Magazine for Nov., 


1912, states that the thermic losses of a compressed gas may be 
6 


82 AIR COMPRESSION AND TRANSMISSION 


expressed by the contraction which it undergoes in cooling. In 
place of the larger volume of power medium which leaves the com- 
pressor, a diminished volume only at the same pressure reaches the 
destination. As in perfect gases contraction due to cooling is in 
direct proportion to the absolute temperature, the fraction formed 
by taking the absolute temperature of the atmosphere as the numera- 
tor and the absolute temperature of the air or gas leaving the 
compressor as the denominator, might be taken as a fair expression 
of the losses caused by the unutilized heating of the gas or air in the 
compressor. 

As this loss would not occur if the gases were compressed isother- 
mally it is debited entirely to the compressor. He cites in illustra- 
tion a compressor of the piston type of 140,000 cu. ft. per hour 
capacity working against 115 lb. per square inch at go r.p.m., in 
which the temperature leaving the compressor was 197° F. and the 
temperature of the atmosphere 41° F., this gives an ‘‘economic 
f 460+41 __ 501 

400197 657 

Energy Compensation.—If an air compressor is driven direct by a 
steam engine with the steam and air cylinders tandem and one 


efficiency”’ o = 76.1 per cent. 


V (LLL LLL] 


VE LLIILZLIL LLL) 


y | y Z == Z 

BL Ae y 

ey (= 
Steam ‘Air 


Fic. 47.—Direct acting steam compressor. 


common piston-rod as shown in Fig. 47, with the valves arranged to 
give a steam and air card as shown, the greatest force is exerted on 
the piston-rod at the time when the least is required in the air cylin- 
der and when the air cylinder needs the greatest force applied to 
expel the compressed air, the least is being applied in the steam 
cylinder. 

Many ingenious contrivances have been devised for storing the 
excess energy developed in the steam cylinder during the beginning 
of the stroke and drawing on this excess during the last part of the 
stroke. 


EFFICIENCIES AND ENERGY COMPENSATION _ .88 


When a fly-wheel is used it must of necessity be very large in 
order to do this, as the amount of energy that can be stored in the fly- 
wheel will depend upon its weight and speed. 

Hydraulic Compensator.—One form of energy compensator is 
shown in Fig. 48, which represents a sketch of a D’Auria non- 
rotative air compressor. 

The desired result is obtained by using a ‘“‘ hydraulic compensator,” 
which consists of a cylinder A fitted with a plunger B carried by 
the same piston-rod that connects the steam and air piston. The 
ends of the compensator cylinder communicate with each other by 


Seomcesas A Comm nant 


Y au, é 
H gf D 


SESE 


Fic. 48.—D’Auria System of energy compensation. 


means of a loop of pipe c-c—c so constructed as to form a very rigid 
bed-plate for the machine, a very desirable feature, as it helps to 
keep the machine in alignment. The cylinder and pipe are filled 
with water, or any other liquid, leakage being made up through a 
pipe. 

When the compensator is in action, the liquid column contained 
in the compensator is moved reciprocally and as it requires energy 
to start a mass moving and also to stop it after it gets in motion, the 
excess energy of the steam cylinder is used up or rather stored in 
starting the liquid in motion during the first part of the stroke, and 
this excess energy is given back during the last part of the stroke as 
the pistons near the end of their stroke. 

Lever Compensation.—Sometimes two steam air compressors are 
placed side by side and the piston-rods connected by a system of 
levers as shown in Fig. 49, so that the excess energy that is not needed 
in one air cylinder is conveyed by the system of levers to the other 
air compressor and aids that near the end of its stroke. By this ~ 
arrangement one compressor supplements the other. 

Weight Compensation.—A method adopted by the Norwalk Iron 
Works is best shown in their two-stage compressor, driven by a 
tandem compound steam engine, as shown in Fig. 50. 


84 AIR COMPRESSION AND TRANSMISSION 


By arranging air and steam cylinders, as shown, with a common 
piston-rod, an excessively heavy moving piece is secured, which 
requires considerable energy to start in motion and also to bring to 
rest near the end of the stroke. That is, a large share of the energy 


Gx ZZ 


i 
ae 
SAAS 


Fic. 49.—Lever system of energy compensation. 


developed in the steam cylinder during the beginning of the 
stroke is used in starting this heavy piece in motion and the extra 
energy required in the air cylinders during the last part of the 
stroke is taken from this moving mass in bringing it to rest. 


gure “ur by 


SSS OSS has yp Ee 2 el 
ae} Ne 


ve 
S =a fe Ml 
S| ae = as E er 
“ 
Pipl ie : 
SS 


Life i= = = 


== 9_ ira ey 


| a 
aes 


Fic. 50.—Norwalk compressor. 


Straight-line Compressor.—The balance is aided still further 
by a fly-wheel which with shaft and eccentrics is used to operate 
the valves. It is evident that an air compressor which has the steam 
cylinder and the air cylinder on the same piston-rod will apply the 
power in the most direct manner and will involve the simplest 
mechanism in construction. 

This type of compressor (Fig. 51) is usually referred to as a 
straight-line air compressor and is usually equipped with one or 


EFFICIENCIES AND ENERGY COMPENSATION 85 


two fly-wheels to act as energy compensators. Even then it is 
difficult to secure a very good economy, especially with light fly- 
wheels. In order to secure maximum economy of steam an early 
cut-off is desirable, but if no fly-wheels are used this cannot be 
obtained, and it is necessary to admit boiler steam for almost the 
entire stroke. 

The air compressor used by the Westinghouse Air Brake Company 
in their familiar system of train brakes is of this type. It is admira- 
bly suited for this purpose because of its simplicity and the fact 


Fic. 51.—Straight-line air compressor. 


that it does the most work when the engine is at rest or using only 
a portion of its steam, and for this reason it utilizes steam that 
might otherwise escape out of the safety valve. 

Many efforts have been made to equalize the steam power and 
air resistance by using a crank shaft and placing the crank pins 
of the steam and air-connecting rods at an angle with each other 
so that the greatest force would be exerted in the steam cylinder 
at the time the greatest resistance was being encountered in the 
air cylinder. The same thing may also be accomplished by placing 
the cylinders at an angle with each other. Various compressors 
have been built on this principle, the angle between the cylinders 
varying in different designs, being in some 45 degrees, in others 
go degrees, and in still others 135 degrees. The best results, how- 
ever, have been secured with an angle of go degrees. 

This arrangement has been adopted by some manufacturers 
of compressors for refrigerating plants, but has not been used by 
manufacturers of air compressors to any extent. Fig. 52 may make 


86 AIR COMPRESSION AND TRANSMISSION 


this clearer, with the horizontal cylinder for steam and the vertical 
one for the ammonia compressor. When the steam piston is at 
dead center the air piston has completed about half its stroke, 
and the high steam pressure admitted to the steam cylinder during 
the first part of the stroke will be available for moving the compressor 
piston through the last half of its stroke when the greatest resistance 
is encountered. As the steam piston is completing the last half 
of its stroke, the compressor piston starts down compressing a new 
supply of free air on its lower side if of the double acting type, and 
as the work of the first half of the stroke of the air piston is com- 


CHEAT TEC EOL CUCE CAE COLEEOEEGy 


Fic. 52.—Horizontal-vertical arrangement of cylinders. 


paratively slight, the pressure in the steam cylinder can be reduced 
for the last half of its stroke, giving both economy of steam and 
uniformity of speed. 

Duplex Compressor.— More frequently this result is accomplished 
by placing the two cylinders in a horizontal plane with the crank 
pins at an angle of 90 degrees as shown by Fig. 53. This arrange- 
ment is frequently adopted when air compressors are driven by 
gas engines and if an air compressor is driven by a belt the compressor 
will operate much more evenly and hence with a more uniform pull 
on the belt if two or more cylinders are used with the crank pins 
of each placed at an angle with each other. 

The “duplex air compressor” is designed on this plan with two 
cylinders side by side, the crank pins for the two compressors 
being at an angle of 90 degrees with each other. The motive 


EFFICIENCIES AND ENERGY COMPENSATION 87 


power may be either belt, electric motor or steam engine. If the 
latter, it is not uncommon to place the steam cylinders tandem 
with the air cylinders, using a common piston rod, as shown in 
Fig, 54. 


NST SY NL__ift 
eS TM Vy 


—Z 
J 


a} N = sl | Dy 


Fic. 54.—Duplex steam-driven compressor. 


The steam cylinders may be either cylinders of a cross-compound 
engine, or two separate simple steam engines. Similarly, the two 
air cylinders may be cylinders of two separate air compressors or 


88 AIR COMPRESSION AND TRANSMISSION 


cylinders of a two-stage compressor. The name ‘“Duplex”’ is 
applied to any of these designs. 

Figure 55 shows a sketch of the arrangement of cylinders for 
a two-stage duplex compressor driven by a cross-compound steam 
engine. 

A little study of these sketches will make it clear that with such 
a duplex arrangement when the greatest power is developed in 
one steam cylinder, this excess power can be utilized by means 
of the common crank shaft in overcoming the maximum resistance 
that is being encountered in the other air cylinder. With the cranks 


ap 


Fic. 55.—Duplex cross compound steam, two-stage air compressor. 


go degrees apart there is little difficulty in starting, even if compound 
steam cylinders are used, for if the compressor would stop with the 
high-pressure cylinder at dead center, live steam may be admitted 
to the low-pressure cylinder by means of a by-pass. 

Commercially, the duplex compressor appeals to the trade in 
that one side or half of the machine may be furnished with fly- 
wheel and out-board bearing designed for a complete machine, 
and as the demand for compressed air increases, the output may 
be increased by installing the remaining side of the machine. 

The belt compressor is probably the best type for small capacities 
when it can be used conveniently, as is the case in a great many 
factories, for the losses in a steam cylinder, especially of small 
power, are excessive as compared with the loss of power due to 
belt transmission. 


Cit’ ili Re EX 


MULTI-STAGE COMPRESSION 


It was pointed out in Chapter VII that it was not advisable to 
attempt compression above 80 lb. per square inch in a single cylinder 
because of the loss of energy and danger of explosion due to the re- 
sulting high temperatures. 7 

It frequently happens, however, that pressures much higher than 
this are demanded for commercial purposes, and in order to satisfy 
this demand, avoid the danger just referred to, and reduce the losses 


MEREBEU AS ses 


SSS SSS 


b 
f 
q 
i 


f es 
je OPO La 
=e SESE EELS ERASE LUE UNEENANUNENENENESSND e 


Fic. 56.—Saving due to multi-stage compression. 


due to adiabatic compression, engineers have adopted a multi-stage 
system of compression; compressing the air partly in one cylinder, 
passing it through an intercooler where its temperature and volume 
are reduced, then compressing it still further in a second cylinder, 
and, if the pressures required are high, this compressed air is passed to 
a second intercooler, thence to a third cylinder and in some cases a 

89 


90 AIR COMPRESSION AND TRANSMISSION 


third intercooler and a fourth cylinder are required to secure the 
desired compression pressure economically. 

Advantage of Multi-stage Compression.—The advantages of 
this system of compression more than offset the extra expense in 
constructing the compressor. The saving in power required may be 
illustrated by Fig. 56, where a—d represents the adiabatic line from 
atmospheric pressure to the required receiver pressure, a—c an iso- 
thermal line between the same pressures. The shaded area repre- 
sents the total work of compression in the four cylinders, the differ- 
ence between this area and the area abde representing the saving in 
power due to the multi-stage system of compression. afne repre- 
sents the work done in the first cylinder, fn the volume occupied by 
the air as it leaves this cylinder. In the intercooler the temperature 
of the air, if this part of the apparatus is properly designed, will be 
reduced to the inlet temperature, and in consequence the volume will 
be reduced from fx to on. Compression in the second cylinder will 
raise the pressure to g and reduce the volume of the compressed air to 
gm. In the second intercooler the volume will be reduced as the 
temperature is reduced to the inlet temperature from gm to pm, and 
so.on. ‘This secures a compression that requires a smaller expendi- 
ture of energy than adiabatic compression, giving results that com- 
pare very favorably with the ideal isothermal compression without 
serious difficulty. 

Pressures Used for Various Stages.—Of course this arrangement 
increases the first cost of the compressor and for that reason the ad- 
visability of installing multi-stage compression will depend upon the 
pressure required. Some authorities recommend two-stage com- 
pression for pressures as low as 50 lb., but this practice is unusual. 
It is certain, however, that for pressures from 80 to 500 lb. the two- 
stage compressor should be used; for pressures from 500 to 1,000 lb. 
the three-stage, and for pressure between 1,000 and 3,000 lb. the four- 
stage compressor. 

Intercoolers.—To secure best results care should be taken to see 
that the intercooler between the different cylinders reduces the 
temperature of the air as nearly as possible to that of the air at the 
compressor inlet. As it is important that the flow of air through 
the intercooler should be as low as possible, it is desirable to reduce 
the pulsating effect of the discharge of partially compressed air to 
the intercooler. This is usually accomplished by using large ports 
and passages. 

The larger the volume of the intercooler, the more time for the 


MULTI-STAGE COMPRESSION 


Fic. 57.—Intercooler for duplex two-stage compressor. 


Wa ter Outlet 


> 
KS 


as See SS | See 
Uy N= 
UJ 
4 
U 

a 


. 
eet oe ak 
id Site esses SY 


eee sf 
(e ~] 

LO: meas ee ‘Of 
$22 Yor 
Ogee Ss (| 
wz J 


cc OL Ne 


(a hig a\'=| 
YTZZALLIP OTP LIZZ ne 


s 


ZZA Rae. 
ML 
T 
| y 
i 


Cima waa waar ar ae 


Re 


RS 


ll Robeech 


ZA U 
— if N MY] 
eS 
<a Ss 
a 


ee ey 


howe 
‘AMIN KADY 
SS Pe rae \ Z) ' 
AN 7 
is 
H Ly 
4 
i wy wanes 
1 GN NOE 
f Zia tH Tomi ah = 
‘a 
HAT SS 
Dp in 
az eee: 
Ac LX ‘ 


Sy H 
we ees Jacket Pipe- 
? Air Discharge///////7 


Fic. 58.—Intercooler for tandem two-stage compressor, 


92 AIR COMPRESSION AND TRANSMISSION 


compressed air to cool; for this reason ‘‘receiver intercoolers,” as 
they are called, are more efficient than those of small volumetric 
capacity. 


Fic. 59.—Nordberg intercooler. 


Types of Intercoolers.—Figs. 57, 58 and 59 show various types of 
intercoolers. The horizontal type is more frequently used because 
of its greater adaptability to 
compressor construction. 

In accordance with the 
fundamental principles of 
economical transference of heat, 
it is customary in better types 
of intercoolers to have the cir- 
culation of the water opposite 
in direction to that of the air 
being cooled, and also to have 
ond i the air broken up into as fine 
Ce ir L—-  streams as possible. 

fi ¥ The tubes of intercoolers are 
usually of iron unless the 

character of the water used is 
bad. In this case the tubes 
may be galvanized or, if the 
water is salt or contains mate- 
rials having a corrosive effect on iron, brass or copper tubes are used. 

In cooling the air, moisture is frequently deposited, and provision 


Fic. 60.—Intercooler with separator. 


MULTI-STAGE COMPRESSION 93 


is made to remove this by traps or separators, as shown in Fig. 60. 

Perfect intercooling implies that the temperature of the partially 
compressed air leaving the intercooler shall be as low as the atmos- 
phere. This naturally requires different ratios of cooling surface to 
cubic feet capacity for different water temperatures. 

Cooling Surface and Capacity.— Mr. F. V. D. Longacre gives two 
charts covering this matter, shown in Figs. 61 and 62. The first 
shows the intercooler surfacer equired for various water tempera- 
tures to secure perfect intercooling for two-stage compression to 100 
Ib. discharge pressure at sea-level; and the second shows the amount 
of water required to secure perfect intercooling for this pressure if 
the cylinder jackets and intercooler are in series, also the amount of 
water required with a separate jacket, and when the low-pressure 
and high-pressure jackets are connected in series. 

Intercooler Pressure.—In considering multi-stage compression, 
it is necessary to determine the proper intercooler pressure to secure 
the most economical results. 

It was shown in Chapter IV that the area of an ideal indicator 
diagram disregarding clearance could be expressed as 


nN Po na 
—"— 144 paVe| (22) -| ft-lb. 


and if Va represents the capacity of the machine in free air per 
minute, the horse-power required, disregarding friction and other 


losses, will be: 
144 Po 
33000 rane aVa (ere -:| 


If pp represents the intercooler pressure, this would represent the 
horse-power required to operate the low-pressure cylinder, and if 
the discharge pressure from the high-pressure cylinder be represented 
by pa the horse-power required to operate this cylinder would be: 


144 n pa — 
33000 pave (22 | 


but if perfect intercooling were secured, poVc would equal paVa 
and the total horse-power required to operate both cylinders of 
the two-stage compressor would be cee 


Petr nt paVe| (22) "= + (2) "* =. 


94 AIR COMPRESSION AND TRANSMISSION 


E 

> 

fe 

= 

S 20 Curve showing intercooler surtace| 

4 10 100 |b. discharge pressure at sea 

c- ae level to obtain pertect intercooling. ee eae 
22) 

a= |b 


z SCS et eee 
£ ad 
= [so one ase ae 
= Tee eee a 
Z, Bae eee Bere 
‘ a [ES ete hit] | a 
t ea [9 5] a a a 
8°9 5 45 50 


35 
Dereace ieniperarre pee Air Peand and Water enone 


Fic. 61.—Intercooler surface required. 


IS 
cS 


BS 
i Se | 


D 
ro 


Gallons of Water per Hour 


Volume of water Sanilgd 
for perfect intercoo 

100 Ib. disch. pressure 

sea level operation Wath 
and without cylinder 

Jacket water. 


ay, on 


Fala 

ELE 
280 

Siew 
=o an ea 
v 260 
77 a 
(allel a NR ESE 
sca) 
tp 
Coy ye 
lay Ae 
sae 

z 

VA 

ie 

a 

Es 


BOP ROES 
eee ine | 
SSE 


120 


VALS 
PAV 
BBVA 


90 
ewetee rCiaiestern beneee Fahrenert 


Fic. 62.—Water required for intercooling. 


MULTI-STAGE COMPRESSION 95 


This expression will be a minimum when the part within the 
brackets is a minimum. As fa, or the inlet pressure, and pa, the 
discharge from the high-pressure cylinder, are fixed, the only 
variable is the intercooler pressure, or po. 


(jak n—1 
Differentiating eo) Oe ian cp ies | with respect to pp and 


equating to zero will give the requisite condition of proper inter- 
cooler pressure for minimum expenditure of energy. 


n—1 in n—1 


1h el fh —1 I Aba AP === == 1 
Pay 3 ee + Se, Ci =o 
Raa 
Laas’ a =— in I TNS = 
: Did ae Doak ee Ts - py 
HES 
ny pe n b—n 1—n 
Pe = a5 a aSpE De wD ee 
po=NV Paba 


That is, for two-stage compression the most economical expendi- 
ture of energy is secured when the intercooler pressure is the square 
root of the product of the given suction and discharge pressure of 
the machine. As perfect intercooling is assumed, paVa=poVo 
and the areas of the two cards must be equal, that is, the most 
economical results are secured when the work of compression is 
divided equally between the two cylinders. 

Let Fig. 63 represent an ideal card of a three-stage air compressor 
without clearance, in which p3 represents the pressure in the first 
intercooler, and 4 the pressure in the second intercooler. These 
first two stages may be considered as two-stage compressors between 
bp; and p,4 in which, for the most economical results, 


b3s=WV pips 


and in the same way and for the same reason, 


pi=V Papo 
from which 

p3= V obo 
and 

pa=V pips? 


The effect of clearance on the above discussion can be shown 
by referring to Fig. 64, showing cards for a two-stage compressor 


96 AIR COMPRESSION AND TRANSMISSION 


with clearance. The area showing the work done is AJKLSTZ, 
which may be considered as AJNF+KLEN—ZSEF, This will 
evidently be a minimum when the expressions for these areas are 
a minimum, but as the expression for ZSEF does not contain the 


Pon a gs eae 


me--—---——-—— — --— -— - -— - 


| Lo 
o 
NY, Se ee ee eee 


0 


Fic. 63.—Proper receiver pressure for multi-stage compression. 


variable pz, this term will drop out in differentiating and, as a 
result, it will follow that the intercooler pressure giving the most 
economical result will be with clearance as without clearance. 


Px = iy 


V 


Fic. 64.—Effect of clearance on receiver pressure. 


The same method will show that most economical receiver pres- 
sures for a three-stage compressor are: 


p3= V pbs and p4= V pipe? 


MULTI-STAGE COMPRESSION 97 


when clearance is considered as when clearance is omitted in the 
discussion. 


Effect of Clearance on Volumetric Efficiency.—It was pointed 
out in Chapter VIII that the real volumetric efficiency of an air 
compressor could be expressed as 


re helmet] 


Figure 64 has assumed the clearance in the various cylinders 
to be proportional, that is, the ratio of clearance volume to piston 
displacement in each cylinder was such that clearance lines of each 
cylinder unite to form a continuous expansion line. 

If C represent the clearance of the low pressure cylinder and po 
represent the intercooler pressure, pa the suction pressure and pe the 
discharge pressure from the high-pressure cylinder, then the real 
volumetric efficiency of a two-stage air compressor may be ex- 


pressed 
I Bpe 4 pi oe Po a | 
re Batee| eI] 


but as Po=N Pabey this may be written 

Tam pi'| | (Patbe?\ > |) 
Tr fe: el ( Pa. es 
Tom Pr}, _¢| (Pe) on— || 
arena aa} 


In the same way, the true volumetric efficiency of a three-stage 
air compressor may be expressed as 


or 


Tam Ps is| (2 nor] 
16s Pam | - Pa J 
in which fe is the discharge pressure from the last cylinder and pa 
the suction pressure of the low-pressure cylinder. 

Figure 45 shows graphically the effect on volumetric efficiency 
of compressing by stages and the resulting advantage in capacity. 


CHAPTER X 


DETAILS OF PISTON AIR COMPRESSORS 


Classification of Valves.—Most of the various types of inlet 
valves for piston air compressors may be divided into two general 
classes: first, those which are automatically opened by atmospheric 
pressure and closed by means of their own inertia or weight, by 
springs, or by air pressure; and second, those which are opened 
and closed by direct and positive mechanical connection with the 
crank-shaft or some other moving part of the machine. Each of 
these classes include many forms of valve design. 

Valves of the first class are entirely automatic in their action, 
their opening and closing points depending entirely upon the con- 
ditions of pressure within the cylinder. However, they have 
certain advantages which will be considered later. Valves of the 
second class, with one or two exceptions, have their points of opening 
and closing fixed without regard to changes in operating conditions, 
and the present tendency among designers and manufacturers 
seems to be toward valves of this class. 

Mechanical Valves.—Nothing can be superior to mechanically 
operated valves when properly adjusted to operating conditions, 
as by their aid several of the losses of air compression have been 
reduced to a minimum. 

On the other hand, faulty adjustment of valves, sometimes 
combined with improper design, renders them extremely low in 
both efficiency and capacity. 


Inlet Valve Setting.—If inlet valves are so set that they open 


almost exactly when the piston is at the end of its stroke, the card 
will indicate absolutely no clearance at either end of the cylinder, 
the clearance air being exhausted into the intake. If the inlet 
valve closes slightly before the piston reaches the end of its suction 
stroke, the volumetric efficiency is also reduced. 

In case the inlet valves are so constructed that they cannot open 
until the clearance air has been expanded to atmospheric pressure, 
the only loss due to this clearance is one of capacity, which may be 
Overcome by an increase in size or speed of piston. If, however, 

98 


Pt yer 


DETAILS OF PISTON AIR COMPRESSORS wh 


the inlet valve opens when the piston is in its extreme position, 
the clearance air is exhausted through the intake, making a direct 
loss of power as well as of capacity. 

Figure 65 is reproduced from a card taken from a machine in 
which the inlet valves were set to open when the piston was exactly 
at the end of its stroke. In most of these cases the exhaust through 
the intake is sufficient to cause considerable noise. Figure 66 shows 
a card from the same machine with the valves set properly for their 
particular pressure. This change in the time of opening the inlet 


pm. 


Fic. 65.—Mechanically Fic. 66.—Mechanically oper- 
operated inlet valve opened at ated inlet valve properly set. 
end of stroke. 


valve has not effected the volume of air discharged, but the power 
required to operate the compressor has been considerably reduced 
and the machine will run more smoothly with less shock to the 
moving parts at the end of the stroke. 

Effect of Changing Discharge Pressure.—lIf the pressure of dis- 
charge is now increased, the former troubles appear again, resulting 
in a card shown by Fig. 67. If this pressure is to be maintained con- 


Z)_4 


Fic. 67.—Effect of increasing Fic. 68.—Effect of decreas- 
discharge pressure. ing discharge pressure. 


tinuously, the inlet valve will have to be adjusted to open a little later 
in order to give the best results. 

In the same way, if for any reason the discharge pressure should 
be reduced after the valves have been set correctly, the indicator 
card will resemble Fig. 68, and if the compressor is to operate con- 


” 


100 AIR COMPRESSION AND TRANSMISSION 


tinuously at this lower pressure the inlet valve will have to be ad- 
justed to open a little earlier. 

Figures 69 and 7o show indicator cards from improperly set 
mechanically operated discharge valves with the defect indicated 
under each. 

Every machine with mechanically operated valves should be 
carefully examined to determine whether they operate at the correct 


Fic. 69.—Mechanically Fic. 70.—Mechanically 
operated discharge valve operated discharge valve 
opening too early. opening too late. 


time, andif not, should be so adjusted in order to raise the efficiency 
of operation. 

The principal disadvantages of the mechancally operated valves 
are the increased cost and the extra attention required to keep the 
valves set properly. 

Automatic Valves.—In compressors using automatic valves, how- 
ever, there is no necessity of timing the valves to suit changes of 
pressure, as the operation of the valves is controlled entirely by the 
conditions or pressure within and without the cylinder. 

It must be remembered that the majority of air compressors 
operate at a fixed pressure of discharge and after the valve is once set 
for this discharge, there is no further need of changing it. 

Both types of valves have advantages and disadvantages peculiar 
to each and a choice of valves should not be made in any important 
instance without a thorough investigation of all the variable factors. 
involved. 

Valve Area.—A very important matter to be considered is the 
inlet valve area or port opening required for the proper action of a 
machine. Asin other points of design, it is necessary to compromise 
the desired ends, for the larger the inlet valve the less will be the 
water-jacketed cylinder surface, and as both are desirable it is im- 
possible to give absolute ratios of inlet areas to cylinder sizes. Some 
designers make inlet areas 5 per cent. of the piston area, and other 


DETAILS OF PISTON AIR COMPRESSORS 101 


designers use as high as 14 per cent. asthe ratio. The design of the 
valve, the cylinder proportions, and the speed of the machine, all 
have an influence in determining this point. 

The following data is given by the chief draftsman of a large com- 
pressor company as the practice of that company resulting from an 
experience of many years: 

“Roughly speaking,” 5,000 ft. per minute for the velocity of the 
air through the valve gives good results. This being the case, a 
slow-running machine would require a smaller valve than a high- 
speed compressor with a ‘piston-inlet’ valve, having a piston speed 
of from 300 to 350 ft. per minute, the inlet area is from 5 to 6 per 
cent. of the piston area. On large compressors with a piston speed 
of from 500 to 600 ft. per minute, the valve area ranges from 6 1/2 
to 7 per cent: of the piston area. 

The discharge valves which are of the poppet type are from ro 
to 12 per cent. of the piston area. 

On machines having both inlet and discharge valves of the poppet 
type, the ratio should be about 12 per cent. for machines of that 
speed. For piston speeds not exceeding 4oo ft. per minute it is 
probable that ro per cent. is enough. 

The area of the discharge valve should not be less than that of the 
inlet, for although the volume of discharge is less than the volume of 
admission, this discharge must take place in a considerably shorter 
space of time. 

Forms of Poppet Valves.—Probably the automatic poppet valve 
is the most common form of valve in use. A few designs are shown 
in Figs. 71 and 72. 


Fic. 71.—Air inlet valve. 


The principal difficulty to guard against in the design of an auto- 
matic valve is to avoid the possibility of the valve itself being 
drawn into the cylinder with the in-rushing air. This may happen 
through the breaking of the spring and disastrous results frequently 
happen, for on the return stroke of the piston, the cylinder head, 
or the piston, or some other part of the apparatus is sure to suffer. 

Figure 73 illustrates a peculiar valve designed for a single-acting 
compressor, 7.¢., a type of piston compressor in which the air com- 


102 AIR COMPRESSION AND TRANSMISSION 


pression takes place on only one side of the piston instead of both as 
is usually the case. JB is the inlet valve which is located in the center 


Fic. 72.—Air discharge valve. 


of the piston and is held on its seat by the spring D. The discharge 
valve A is a radical departure from the older designs of compressor 


Fic. 73.—Valve in cylinder fee 


valves, being a flat disc covering the entire area of the cylinder and 
held in its seat by a guide and spring. 


AS 
AA rz raedsaiteg ae 
Ee 
a = ——ews 

AL OH a A fa 


om | 
zee ‘* 4 
is Ae 
Se iA eek ay 
S 


Le ed 


N 
Vy 


GPR 


iN 
yee 


LL 


Fic. 74.—Piston inlet valve. 


Its face and the face of the piston are perfectly flat, so that the 
piston may strike the valve and deliver all the air with no clearance 


DETAILS OF PISTON AIR COMPRESSORS 103 


space to reduce its capacity. A large area of discharge is obtained 
by a very small movement of the valve and no pounding is made 
by its action, for the compressor is of the straight-line type and the 
compression in the steam cylinder acts as a cushion to relieve the 
pounding that might otherwise occur with no clearance. 

Piston-inlet Valves.—One of the most interesting forms of inlet 
valve is the piston-inlet valve, as manufactured by the Ingersoll- 
Rand Company, a sketch of which is shown in Fig. 74. By 
this arrangement the entering air comes in through the tube 4, 
which projects through the head end of the cylinder. The air 
passes through this to the center of the piston which is hollow. 
Communication is obtained from this hollow piston to the cylinder 
through the ring-shaped valves B, which are made of open-hearth 
steel in one piece without a weld. These valves have a movement 
of about 1/4 in., and are held in place by pins which are set in slots 
in the valve. 

The two inlet valves B and the tube A are carried back and forth 
with the piston. The valve on that face of the piston which is 
towaid the right is closed as the piston moves to the right, while 
that on the left side is open to admit a fresh supply of air to the 
left side of the piston while air is being compressed on the right 
side. - 

When the piston reaches the end of its stroke, the inlet valve 
closes because of its own inertia and as the piston starts on the return 
stroke the valve that was formerly closed is now left behind for about 
1/4 in. of the piston travel and remains open during the entire 
stroke. When the valves are closed, their face is almost flush with 
the piston face, thus reducing the clearance space to minimum. 

There are no springs in the construction of this valve, and it has 
been found to work equally well with slow or high speed. These 
valves are guaranteed by the company for five years. 

Discharge valves are shown at H. These conduct the com- 
pressed air to the discharge pipe F. 

All these types of automatically operated valves have the ad- 
vantage that they adjust themselves to meet varying changes in air 
pressure automatically. 

Semi-mechanical Valves.—There are several types of semi- 
mechanically operated valves on the market. Some of these con- 
sist essentially of an arrangement of levers to remove the action of 
the spring on the inlet valves during admission, permitting the valve 
to open instantly and freely and remain open without any clattering 


104 AIR COMPRESSION AND TRANSMISSION 


until the end of the stroke, when the spring tension is permitted to 
act on the valve and close it. 

Attention has already been called to the fact that many auto- 
matic valves close before the end of the stroke, due to the fact that the 
piston is rapidly reducing its speed at that time. If the inlet valve 


Fic. 75.—Mechanical valve of Corliss type. 


closes before the end of the stroke, the volumetric efficiency is natur- 
ally reduced, and on this account the mechanically operated inlet 
valve is preferred by many engineers. 

In addition to the mechanical operation of poppet spring valves 
just mentioned, some air compressors are equipped with valves 


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Fic. 76.—Southwork blowing engine valve. 


which resemble in action and appearance Corliss steam valves. 
Fig. 75 gives an illustration of this form of valve, which is opened 
and closed by a rotating motion, given to them by levers from a 
wrist plate or eccentric. 

This type of valve is sometimes used for the discharge valve on 
compressors, but cannot be operated successfully for very high 


DETAILS OF PISTON AIR COMPRESSORS 105 


pressures because the clearance is made excessive. Mechanically 
operated valves are usually used on large blowing engines for blast 
furnaces. One type is shown in Figs. 76 and 77. In this type of 
compressor, it is desirable to secure large, free opening for suction, 
and one of the latest designs consists of a large cylinder on the out- 
side of the compressing cylinder which reciprocates back and forth, 
and in so doing opens large slots at the end of the cylinder, giving 
very free opening for inlet. 


a 


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vi 
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Fic. 77.—Kennedy blowing engine valve. 


Regulators, Unloading Devices, Etc. 


It is often essential that the pressure of air in an air receiver be 
kept constantly at a fixed point and as the number of tools using 
air at the same time will vary in any installation, some automatic 
device must be used so that the_compressor will be furnishing air 
when needed and when no air is needed this device must prevent any 
unnecessary work being done at the compressor. There will be 
times when every tool that is taking air from the receiver will be in 
operation and the compressor must have a capacity sufficient for 
such occasions; and again there will be times when none of the tools 
are in operation and the work at the time being done at the com- 
pressor would be in excess of the needs if some automatic system 
of regulation is not used. 

Belt Regulator.—Probably the simplest form of regulator is the 
one that is often used on belt-driven air compressors. ‘This consists 
of a belt-shifting device so arranged that when the pressure gets 
above the desired point the belt is shifted off the compressor wheel 
and onto a loose pulley. When the pressure falls below this fixed 
point the belt is shifted back again, and the compressor is thus 


106 AIR COMPRESSION AND TRANSMISSION 


automatically started and stopped to suit the changing amount 
of compressed air that is needed. 

Westinghouse Governor.—Fig. 78 shows a sketch of the governor 
used on the Westinghouse air brake. It consists of a piston A 
moving in a cylinder and directly connected to the steam valve C 
which supplies steam to the air compressor, or air pump as it is 
more commonly called. A spring, D, helps to hold this up and 
hence keep the steam valve open. Pipe £ leads from the air reser- 


%, D> 


TAT TTD ay 
praia asst 
Y 


WH: 
‘t 
su w/ 


| 
: 


Kok 
Lad 


Uy 
\ 
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MT 


\ 
Be, 


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K Vlsse 

WN 

NU 
Yi 


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VAL ey) 


Zbl 


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2 


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Z 
% 


Fic. 78.—Westinghouse governor. 


voir to the governor. Communication between E and the cylinder 
above A is closed by a needle valve Ff, which is held on its seat by 
the governor spring K. When the air pressure in the main reservoir 
gets up to its maximum, the pressure in £ is sufficient to raise the 
small piston H against the governor spring K, lift Ff from its seat 
and allow the air to press on A and thus close the steam valve and 
stop the air pump. A small opening, L, allows the air above A to 
escape gradually into the atmosphere. As air is used in releasing 


DETAILS OF PISTON AIR COMPRESSORS 107 


the brakes, the pressure in the reservoir will be reduced, and when 
this happens, spring K can overcome the air pressure and seat F and 
spring D will then raise piston A, open the steam valve C and start 
the pump. 

The governor used for controlling the compressor for electric- 
driven air-brake systems consists of an ordinary Bourbon pressure 
gage with a special needle or hand, which upon coming in contact 
with a stud at the position of minimum pressure causes an electric 
current to flow through a magnet coil. This coil operates a plunger 
to which the contact pieces for the motor circuit are attached and 


MeL 


Fic. 79.—Belt regulator. 


in this way the circuit is closed and the motor started. As soon 
as the air pressure reaches the desired maximum the gage hand 
strikes another stud, causing current to pass through a second 
solenoid magnet which pulls the plunger referred to in the opposite 
direction and stops the compressor motor. 

By this mechanism it is possible to get a close margin between 
the maximum and minimum pressure. This margin can be changed 
by moving the studs. 

Figure 79 shows a form of regulator for belt-driven compressors 
~which stops the compressor when the desired maximum pressure 


108 AIR COMPRESSION AND TRANSMISSION 


is reached. When the desired upper limit is reached the belt is 
shifted from the tight to the loose pulley. 

Norwalk Regulator.—One of the simplest forms of regulators 
for steam-driven air compressors is the one made by the Norwalk 
Iron Works shown in Fig. 80. It consists of a balanced steam 
valve A placed in the steam-pipe near the steam cylinder and con- 
trolled by the air pressure in the receiver. A small cylinder B 

contains a piston connected with the 
opp balanced valve A by the stem C. Above 
s this small piston is a stop screw D pro- 
jecting above the cylinder head for regulat- 
ing the lift of the piston. The air from the 
receiver is led through a small safety valve 
E which regulates the pressure at which air 
can escape into the cylinder B to move the 
piston. Above the disc of the small safety 
y valve is a spring whose tension is regulated 
i by ascrew F allowing the pressure at which 
air is permitted to enter cylinder B to be 
changed at will. The air passes into cylin- 
der B below the piston and if no escape 
were provided would drive the piston to 
the top of B. To regulate this a very fine 
slot is cut in the side of the small cylinder. When the piston 
rises it uncovers this slot and thus furnishes an escape for the 
air which is passing the safety valve. If only a little air enters 
then a small part of the slot will accommodate it and the 
piston will take a low position. With more air escaping the piston 
will rise higher and uncover more of the slot, thus providing a 
larger opening for its exit. As the slot is very fine, a very little 
difference in the quantity of air will cause the piston to assume a 
high or low position. After the small safety valve begins to blow 
an almost insensible increase of pressure in the reservoir will furnish 
enough more air to carry the piston to the top of the cylinder. 
Thus any degree of regulation is obtained by a very little difference 
of pressure, as the air which works on the piston in the small cylinder 
has only to perform the work of lifting the piston and valve suff- 
ciently to uncover enough of the slot so that it can escape; its pressure 
is very slight. 

The piston is fitted loosely and the whole apparatus moves as 

nearly without friction as can be imagined. 


aaear 
Hy 
sere 
Hj 


TI tis 


Fic. 80.—Norwalk 
governor. 


DETAILS OF PISTON AIR COMPRESSORS 109 


When this regulator is applied to compressors having a single 
steam cylinder, it is possible for the valve to be carried so high as to 
cut off all steam and to stop the engine on the center. This would 
be objectionable. To obviate this, there is placed on the top of 
the small cylinder a screw stop which can be set to prevent the 
closing of the steam valve more than is sufficient to run the engine 
at the slowest possible speed. 

Combined Governor and Regulator.—Another combined speed 
and air-pressure governor is shown in Fig. 81. This not only 


Breas 


: arate 


sean LN aN Nalalol ahaha # 


ca ae en 


aN 
| 
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De PE LP PT A wr 


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LZ 

<2 
<u 
V7 


ISSISSEB SS 


Fic. 81.—Clayton governor. 


performs the functions of an air governor, but also prevents the 
compressor from operating at an injurious speed should a sudden 
drop in the air pressure produce a greater demand upon the com- 
pressor than its highest reasonable speed can supply. It consists 
simply of an adjustable stop attached to an ordinary centrifugal 
ball governor. This stop is adjusted to suit varying pressures of 
air in the receiver caused by the varying demands that are made on it. 

Nordberg Governor.—A combined air and speed governor manu- 
factured by the Nordberg Manufacturing Company of Milwaukee, 
Wisconsin, is shown by Fig. 82. In this type of governor the speed 


110 AIR COMPRESSION AND TRANSMISSION 


of the engine is controlled not only by the centrifugal action of the 
governor but also by any variation of the air pressure. The arm A 
controls the point of ‘‘cut-off”’ for the steam cylinder and is operated 
by the movement of the bell-crank C about the fixed point D. The 
rod E controls the bell-crank and is connected to what is called a 
floating lever B. This lever B is con- 
nected with the centrifugal governor FP 
at J, and with a piston which is in com- 
munication with the air pressure at G and 
is held up by a weight ZH. 

It is evident from this arrangement 
that if the point G should remain station- 
ary and the point J should lower, rod E 
will be forced downward and Ato the 
right; also, if point J should remain sta- 
tionary and point G should rise, the same 
movement will occur, and vice versa. 

That is, if the air pressure should rise 
above normal, the engine will have its 
supply of steam per stroke reduced, and 
if the air pressure should fall, the supply 
of steam will be increased; or, if the pres- 
sure of air remains constant, the governor 
will have the same control over the speed 
of the engine that the ordinary centrifugal 
governor has. 

Fic. 82.—Nordberg Unloading Devices.—Sometimes an air 

governor. compressor must be kept running at con- 
/ stant speed, and in order to prevent it 
from doing unnecessary work when the consumption is not equal to the 
capacity of the compressor, a device is used to remove the work or 
load of the air piston and allow it to move back and forth in its cylinder 
without doing any work. ‘These are called unloading devices. Fig. 
83 shows the principle upon which many of them operate. 61 repre- 
sents a valve on the inlet pipe which is closed when the load is to be 
removed, preventing air from entering the cylinder. These unload- 
ing devices are frequently made use of in starting air compressors 
without any load until full speed is reached, when the load is put on 
as desired. 

Clearance Unloader.—One of the most recent unloading devices is 

arranged to vary the clearance on the compressor as the load changes. 


DETAILS OF PISTON ALR COMPRESSORS 111 


This device is illustrated in Fig. 84 and its effect or operation is 
shown by the indicator cards of Fig. 85. 

This automatic clearance controller, as it is called, consists of a 
number of clearance pockets which are thrown automatically into 
communication with the ends of each air cylinder in proper succession, 
this process being controlled by a predetermined variation in re- 
celver pressure. _ 


(Gz ehhh hhh 


TEA 


= (i 


TIZIIMBTIT COR 


Liz 


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‘uc = 
CZ ZZ ZZ 


‘SI 


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2) 
las 


ZZ 


SSS 
---4"Pipe from Receiver 


SISA 


MASS 


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tf 


SSS 


NS 


in 


<< 
<S 


ae) 
ce 


Fic. 83.—Rand imperial unloader. 41, unloader body; 49, unloader cup 
leather; 50, unloader follower; 51, unloader follower screw; 52, unloader cylinder; 
53, unloader cylinder cap screw; 54, unloader regulator cylinder body; 55, unloader 
adjusting plug; 56, unloader valve spring; 57, unloader piston; 58, unloader inlet 
plug; 59, unloader nipple; 60, unloader dirt collector; 61, unloader plunger. 


Regulation is obtained in five stages, viz., full-load; three-quarter- 
load; half-load; quarter-load, and no-load. 

When the compressor is operating at partial capacity, the clear- 
ance space of the compressor is increased and, as a result, its volumet- 
ric efficiency is reduced without changing the suction or discharge 
pressures, or the speed of the compressor. — 

On two-stage compressors these controllers are placed on both 
cylinders, thus maintaining a constant ratio of compression. The 
device is automatic and very satisfactory for use on compressors 
which are motor driven, or must, because of their method of opera- 
tion, be driven at constant speed. 


112 AIR COMPRESSION AND TRANSMISSION 


sey 
= Y 
oe 


Fic. 84.—Clearance unloader. 


Low Pressure Side 


High Pressure Side 
Scale 24 Scale 60 
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Third Step Third Step 
LHP 31.3, H.P Input 97 
Volume 8%, PF 94 % 
~~. MEP 586 MEP 556_-- 
~ M.E.PL84 -7 oe ne 
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Fourth Step Fourth Step 


Fic. 85.—Cards showing clearance unloader. 


CHAPTER XI 


TURBO-COMPRESSORS 


The Engineering Magazine in recent issues has given a series of 
five articles on turbo-blowers and compressors by Franz zur Nedden, 
Superintending Engineer of Weise and Monski, Halle-Saale, Ger- 
many, from which the following material and illustrations have been 
gathered. 

The introduction of the steam turbine as a competitor of the 
reciprocating engine has necessitated a similar change in the design 
and construction of pumps, blowers, and compressors and has natu- 
rally led to the production of turbine machines built for compress- 
ing air or gases. 

The advantages of turbo-compressors, however, are not so apparent 
in small as in large size units, and the high cost of such units in the 
experimental development of this machine has made its introduction 
rather slow. 

The recent development, however, of exhaust steam turbines has 
stimulated the use of turbo-blowers. 

‘“‘Tt is well known that the economy of the steam turbine increases 
directly with its rotative speed, and even electric generators of the 
highest speeds are slow-running machines when compared with the 
steam turbine operating at the number of revolutions required to 
secure the best economy. The turbo-compressor, like the steam 
turbine, becomes more and more economical the faster it runs, and 
is therefore a proper companion of the steam turbine. Speeds of 
4,000 revolutions per minute and even more are not unusual to the 
design of turbo-compressors. 

“Tf large volumes of compressed air are wanted in plants where 
considerable quantities of exhaust steam are available at the same 
time, the coupling of an exhaust-steam turbine with a centrifugal 
compressor becomes an ideal arrangement, and the combination is 
far superior in economy to the piston compressor driven by an 
electric motor or a high-pressure steam engine.”’ 


Design of Turbo-compressors 


“In studying the development of turbo-compressors it is most 
interesting to observe that Rateau and Parsons dealt with the prob- 
lems of design quite differently. 

8 113 


114 AIR COMPRESSION AND TRANSMISSION 


‘“‘Rateau Blower.—Prof. Rateau did not take the structural 
features of his new machine from his steam turbine, but from his 
high-lift turbine pump. His turbo-compressor and turbo-pump are 
so similar that a superficial inspection of the drawing of the two 
~ machines might not reveal the difference (Fig 86). 


Za pe OO OL OO LOE 


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SWS 
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lf orn 
f} \ 


Fic. 86.—Original Rateau turbine blower. 


“The air upon entering the impeller near its nave is seized by 
the impeller blades and thrown outward radially. Its kinetic energy 
due to velocity upon leaving the periphery is changed into pressure 
in the fixed diffusor channel, and being led back toward the center 
the air enters the second impeller to undergo the same process in a 
second stage and so on. Some essential differences in the design 
of details will be taken up later on. 

“The Parsons Blower.—Mr. Parsons, on the other hand, made 
the turbo-blower merely an inversion of his steam turbine. Fig. 87 
shows a unit consisting of a standard Parsons steam turbine (on 


© 
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wv ie z My \ Rspstissssscy SE) 

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ANGE 


L244 


rat 


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LL 


d 


Fic. 87.—Parsons turbine blower with steam turbine. 


the left-hand side) coupled direct with a turbo-blower of the usual 
Parsons type, and delivering 1,600 cu. ft. per minute against a pres- 
sure of 6 to 20 in. of mercury at speeds varying from 2,400 to 3,400 
revolutions. 

“The air is drawn into the chamber B, and conducted into the 
periphery blades of the runner A by fixed guide blades D. The 
following blades are not shown in the section merely for simplicity 
of outline. 


TURBO-COMPRESSORS 115 


“The principal divergence from the Rateau design is that the 
impellers of the Parsons turbo-blower throw the air in an axial direc- 
tion to the next guide apparatus. Parsons undertook to transform 
the kinetic energy of the air as it leaves the impeller blades into 
pressure by simply opposing plain straight blades against its flow. 
The second guide-wheel transmits the air axially to the second 
impeller, which again throws it axially into the third guide-wheel, 
and soon. Fig. 88 shows a developed section through several rows 
of blades. | 

“The excellent reputation of the Parsons machines helped the 
introduction of his turbo-blower, which was rapidly accepted and 


PVE be 


| 


~e 
| | | Guide Me 


POON oe 


Fic.” 88.—Developed section of Parson’s blades. 


put into practical operation. At a time when more than a dozen 
Parsons blowers were in operation or under construction, Rateau 
was still experimenting with his first turbo-blower. Nevertheless, 
Prof. Rateau succeeded in making up this delay and soon advanced 
to the point of combining several of his blowers in series, thus pro- 
ceeding to obtain final pressures of 100 to 150 lb. per square inch. 
The excellent results which he and his assistant, Piof. Armengaud, 
obtained from their high-pressure machines induced even the 
licensees of Parsons steam-turbine patents to secure rights for the 
Rateau turbo-compressors. A careful comparison of both systems 
will disclose seme reasons for the rapid adoption of the Rateau 
system. 

“Cooling Turbo-compressors.—Increase of temperature makes 
special cooling arrangements indispensable, especially with turbo- 
compressors, 7.¢., with machines compressing air to more than 20 
lb. per square inch absolute pressure. Economical cooling becomes 


116 AIR COMPRESSION AND TRANSMISSION 


a vital question in the thermal efficiency of the compressor. On 
this point it decidedly excels the piston compressor, as it is impos- 
sible to cool the air continuously when it is compressed in cylinders. 
(See Figs. 89 and 90.) Here jacket cooling is the most important 
part of the whole cooling system, and the special intermediate 
coolers used between the separate cylinders of compound piston 


Quant. 
ell qiry 


xO LN 
SS 
letotoneteronentansioes 


Aa a 


FRB ae 
£277 
LLL Lace —— 


SOnerIC AIP. a AT A A a 


Lid Atmospheric Alt ———— 
per yee is 


5 10 
Yolume, Cubic Feet per Haur 


Fic. 89.—Diagram three-stage piston compression. 


compressors are generally considered wholly unnecessary in the turbo- 
compressor. In the turbo-machine compression of the air proceeds 
much more gradually, the distance traveled by every particle of 
air is consequently much greater than with piston compressors, and 
the entire area available for the cooling influence of the water is 


BUS 
Va KK Vili 
eee SSS Coble Feet wee 


0 500 ; 1000 1500 
Volume, Cubic Feet per Hour” 


Fic. 90.—Diagram of tubo compression. 


many times as large as that in the piston compressor of equal capac- 
ity; therefore, the air pumped by the tucbo-compressor can be kept 
at nearly constant temperature throughout the operation. Cooling 
arrangements of the counter-current type can easily be used to give 
maximum effectiveness, a condition not readily attained in com- 
pressors of piston type. 


TURBO-COMPRESSORS isles 


COOLING DEVICES 


“The principal differences noticeable between various turbo- 
compressor systems are in their cooling arrangements. The various 
licensees of Prof. Rateau do not use a uniform cooling device. Fig. 
g1 shows one of the first water-cooled Rateau compressors which 
has been successful in practical operation. Fig. 92 shows its internal 
features. Each of the three groups coupled in series contains seven 


oY Ni Nees iN MMos A) 


Yi 
YA ae : 
Al aime 4 


—— ns | 
if = zm 7 ain iff i 
Ue il, — LH Wii i 


Fic. 91.—Water-cooled turbo compressor. 


or eight stages. Each casing is separable horizontally into two parts, 
a form which seems to have become standard for turbo-compressors 
as it has for steam turbines. The cooling water enters the casing 
from below at the highest pressure stage of the group. It passes 
thence to the upper part through copper tubes, shown in Fig. 91, 
goes through a core-hole at the top of the highest stage into the 
upper half of the highest ons but one, and again through copper 


“ HT bn ' 
ae _—————Sa 
i oy 


WT, ———— LLLLLLLLLLLLL LLL Ae 


ee ecg er cnemene es teen ce: PAs0-a=> 
: 


Fic. 92.—Turbo compressor built by Brown, Boveri and Co. 


tubes into the lower half, whence it goes through a core-hole into 
the lower half of the next lower stage, and soon. Later, the copper 
pipes were replaced by small bored holes passing through the 
horizontal joints. (See Fig. 92.) This system has the disadvantage 
that there is no assurance that the water shall completely fill up the 
cooling chamber, as the core-holes are not always at the very highest 
points of the chambers. 

‘““Messrs. Brown, Boveri & Co. avoided this difficulty, and, more- 


118 AIR COMPRESSION AND TRANSMISSION 


over, greatly enlarged the cooling area by casting the vanes hollow, 
thus leading the air back to the center and creating a separate inte- 
rior cooling chamber B, Fig. 93. The water, after filling chamber A, 
flows through the hollow guide vane into chamber B, and thence by 

Pas ek pipe D (which is screwed into 


highest point of chamber A of the 
next stage. Though the excellence of 
this system cannot be denied, it is, on 
the other hand, very expensive, for the 
castings become highly complicated, 
and the foundry work, moreover, must 
be such as to guarantee that all sur- 
faces are absolutely smooth, as the 
elle frictional resistance of air is dependent 
on the roughness of the surface over 
which it passes. The cost of casting 
these casings was about 5 1/4 cents a 
pound in place of 3 1/2 for average casings, and even then this 
famous foundry could not avoid 15 to 20 per cent. waste. 

E “‘C. H. Jaeger of Leipzig wisely separates the casing of his turbo- 
compressors into as many chambers as there are stages, and screws 
these together as shown in Fig. 94. Furthermore, he separates each 


SSSSOQq 


Y 
A 
Y 
WV, 
y 
y 
V) 
mY, 
j 
bod) 


Fic. 93.—Ipmroved cooling sys- 
tem for turbo compression. 


Sen wal NS avtall uu 
Winch aaa wee Ne i ee a a= i ed, 
META iy LLL LL LULL i! UML AA As ze 
‘ oa = ee 
IN = W\ 


ill 
NS WA WA 


Fic. 94.—Water-cooled Jaeger turbo compressor. 


stage into an upper and lower half. Though this system increases 
considerably the number of machined surfaces, it nevertheless 
insures small casings and enables the maker to manufacture the 
single stages on a large scale and to combine out of stock as many as 
are needed for any special requirement. 

“Expansion of Casing.—Another effect of temperature rise, and 
one which acts on the machine, is the expansion of the casing by 
heat. Special preventive measures must therefore be taken to avoid 


TURBO-COMPRESSORS 119 


any alteration in the relative position of the casing and the runners. 
As the casing rests by lateral supports.on the bed-plate, the absolute 
height of its center above that bed-plate will change as soon as 
the casing becomes heated and expands. If the shaft, revolving 
within this casing with a clearance as small as 1/1000 of an inch, is 
supported by bearings which remain practically cool, the distance 
between the center of the shaft 
and the bed-plate will remain 
unchanged. Therefore, the 
centers of the casing and of 
the shaft, which may coincide 
when both are cold, must 
differ as soon as the blower 
comes into operation and the 
temperature of the casing 
rises. The clearances will 
then not only become eccen- 
tric, but very probably the 
shaft and the vanes of the im- 
pellers will come into close 
contact with the fixed parts, 
causing heavy friction, and 
because of the absence of any 
lubricating medium they would very likely seize. 

“Designers of turbo-compressors have overcome the difficulty of 
axial expansion by means already well known in the steam engine 
and the piston compressor. These difficulties are, of course, larger 
with long-extended turbo-compressors than with single-casing turbo- 
blowers. The blower illustrated in Fig. 95 rests with one end only 
on the bed-plate, the other end being free to move or expand. When 
additional stages are required, the design must be altered tc a form 
now quite generally adopted by the licensees both of Prof. Rateau 
and Messrs. C. H. Jaeger & Co. The body is supported only by the 
two terminal covers which carry the bearings. One of the bearings 
is fixed rigidly to the bed-plate, while the other is allowed to slide 
to some extent on the machine rest. (See Fig. 94.) With very long 
blowers, it might perhaps be advisable to support the body by 
lateral feet, the machined surfaces of which might move freely on 
the bed-plate on exactly the same horizontal plane as the axis of the 
blower. In this way radial expansion would not alter the height of 
the geometric center of the casing above the bed-plate. 

‘“fRunners.— Many of the problems that had been solved in design- 
ing steam turbines assisted in the solution of the design of turbo- 
blowers, but although Parsons was able to adapt his steam-turbine 


S Wsr¥% 
N TA 
N= ‘=e 


Neth 
=e 
42 Is 


Wy 


dd 


/ Wi 
Fic. 95.—Jaeger’s turbo-blower. 


120 AIR COMPRESSION AND TRANSMISSION 


runners to his blower, the original designs of Rateau and Jaeger had 
unsymmetrical runners, which had a tendency to become deformed 
under high speed. This difficulty was finally overcome by the use 
- of impellers made with a solid hub with blades and lateral flanks of 
pressed sheet-nickel steel, as shown in Fig. 96. 

“It is not very difficult to insure tightness between the single 
stages of turbo-blowers and turbo-compressors. The pressure 


Fic. 96.—Jaeger’s patent impeller. 


differences against which the clearances have to be kept tight are 
comparatively small, as the delivery pressure is distributed over 
many stages; and, on the other hand, long clearances are secured 
almost automatically as the stages are placed one after another in 
the casing. With air, as with water, the leakage resistance of a 
clearance is greater the longer and narrower it is, but the problem 


Z ) 
iMlalal: 
Fic. 97.—Labyrinth bushing. 


of maintaining tightness against leakage with gaseous media is 
facilitated by the labyrinth effect which is utilized on a large scale 
in the manufacture of steam turbines. 

‘“‘Figure 97 shows such a labyrinth as used by Brown, Boveri & 
Co., the action being briefly described thus: 

“The air passing through the small clearance A expands as it 
enters the following chamber B. By this expansion its pressure is 
greatly decreased, and it traverses the second clearance C at a con- 


TURBO-COMPRESSORS a 


siderably lower pressure than that at which it passed through 4; 
owing to the expansion in the second chamber D, still less pressure 
is left for forcing the air through the third clearance F, and so on, 
and it is therefore impossible for any considerable quantity of air to 
pass the labyrinth. 

“Balancing Axial Thrust.—The only point at which any great loss 
of air occurs is in connection with the usual methods for balancing 
the axial thrust. Special arrangements for this balancing become 
necessary, for, as with centrifugal pumps, the annular area opposite 
the entrance to each impeller is subject to a heavier pressure than 
the entrance itself. And, as with centrifugal pumps, there are sev- 
eral ways of obtaining perfect balance, perhaps the best being that 
illustrated in Fig. 98. 


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Fic. 98.—Turbo blower of 25,000 cu. ft. capacity. 


“Balance by Counter-position.—Here the blower draws the air 
from both sides, and delivers it after both halves of the entire quan- 
tity have been compressed separately in wheels of the same dimen- 
sions through which they pass in opposite directions. Thus, any 
axial thrust arising in a wheel on one side is balanced by an equal 
thrust exerted by a corresponding wheel on the other side. This 
advantage of balancing axial thrust most perfectly, and practically 
without any leakage losses, is paid for, however, in this instance by 
a cumbersome arrangement and poor efficiency. It is obvious that 
the use of double the number of rotating sheaves for compressing 
the same quantity of air doubles also the amount of energy lost by 
frictional resistance. In other words, a blower like that shown in 
Fig. 98 is simply two blowers, each for half the delivery, coupled 
in parallel; and, as the efficiency of all rotating machines drops with 


122 AIR COMPRESSION AND TRANSMISSION 


decreasing delivery, it is clear that the efficiency of these two blowers 
must be lower than if the whole quantity were dealt with in one single 
machine. . 

‘Balancing by Diminishing Back Area.—Another design which is 
standard in the blowers of some of Prof. Rateau’s licensees, for 


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Fic. 100.—Piston-balanced turbo compressor. 


example the Skoda-Werke at Pilsen, and Messrs. Kuehnle, Kopp 
& Kausch, at Frankenthal, is adopted from the well-known Rateau 
turbine pumps. ‘The excess pressure acting on the back of each 


TURBO-COMPRESSORS 123 


impeller can be cancelled by simply diminishing this back area pro 
rata with the increase in pressure. This is done by leaving free an 
annular margin at the periphery of the back of each impeller as 
shown by Fig. 99. The disadvantage of heavy wear and tear, 
which appears when this system is used in turbine pumps, is certainly 
very much less serious when atmospheric air is dealt with. Never- 
theless, there is a certain loss due to the formation of vortices in the 
cells of the impeller where the air is confined by a rotating disc on 
one side and an immovable casing on the other side, and this defect 
is unavoidable in this otherwise excellent balancing system. 

‘“‘Balancing by Balancing Piston.—The design which is now prac- 
tically standard, Fig. 100, is characterized by a continuous flow of 
air through the impellers on one direction, while the consequent 
axial thrust is reduced by other means. 

‘“‘Beyond the last stage there is fixed on the shaft a piston which 
is of the same diameter as the entrance of the impellers and extends 
as far as possible axially. It revolves with a very small clearance in 
a box containing a labyrinth, such as that described above. 

“‘One side of this piston is under the full pressure of the last stage, 
the other side being connected by a large pipe to the entrance cham- 
ber. The effort which the pressures of all the stages exert on the 
piston area is thus equal to the effort which the pressures of the 
single stages altogether exert on the single impellers, and as the two 
forces act in» opposite directions the balance is perfect. Here 
again is another reason for limiting the number of stages that may 
be coupled together in a single casing, for although it is possible 
normally to keep the losses caused by the balancing piston of a 
blower within limits of 1 per cent., this would not be possible if the 
pressure differences were as great as 50 lb. or perhaps 100 lb. per 
square inch. 

“‘Stuffing-boxes.—The problem of a reliable, tight stuffing-box, 
which is so difficult in the design of turbo-pumps, can be most per- 
fectly and easily solved in the turbo-compressor. The first group 
of a compressor, or the whole of any blower which is of the piston- 
balance type just described, needs, we might say, no stuffing-box 
at all, as both free ends of the machine are under atmospheric pres- 
sure, and if air should be sucked in around the shafts, no harm would 
be done, as air is just what the blower requires; only in cases where 
the blower has to deal with very poisonous or valuable gases would 
it be necessary to provide a special packing. In all the following 
casings (that is, in casings containing the high-pressure groups), 
stuffing-boxes may be wholly dispensed with by using a fully capped 
bearing, as in Fig. 92. The probability of water mixing with the 
oil would, of course, forbid this design with any good centrifugal 
pump or water turbine.” 


124 AIR COMPRESSION AND TRANSMISSION 


“Figure ror shows test curves taken during the practical operation 
of one of the latest turbine blowers supplied by C. H. Jaeger & Co. 
to the Grillo Zinc Works, Ltd., Hamborn on the Rhine. The curves 
show the relation between pressure, power absorbed by the blower 
spindle, efficiency, and duty at constant running speed. The pressure, 
horse-power, and efficiency, are ordinates over the respective duties 
as abscisse, in like manner to that followed in making these curves 
for turbine pumps. It is easily seen that over a fairly wide range 
the pressure and efficiency remain nearly constant, and the regulation 
of the capacity can therefore be made by simply throttling down 
the surplus quantity. The speed of revolution need not be altered, 
and alternating-current motors oran exhaust-steam turbine may 
readily be used for driving these blowers in their simplest form. 


m.m. Water Column B.H.P. 
os +H | Ane 
ae a ee ES ho 
100 
2800 / pressuce eae as l 
eee ste 
Hee 
52000 i Sar | | TT 10 a: 
pee -T IA Alea 
ae 1600 80 A ae aS Y 
5 [Le A ipameceae || 508 
200 60 ae" ! Lap. S 
| (S) [ame SU 400 
dear BLL N Md 
g00 “i4q} Lo es 
= eae ne ST | | Nie 
2 
400 520 | EA ree 
5 yre behing the Blower a ee 10 
5 eZ Laessure penis MTT Uh 


0 10 2 30 40 50 6 7 8&8 9% 100 {10 (20 120 40 
-Quantity Sucked 


Fic. to1.—Test curves, Jaeger’s turbo-blower. 


‘Coupling Compressors.—In some cases, as, for instance, in 
blast furnaces, it may be necessary to generate an extra high pres- 
sure for short periods. Such necessities may arise, for example, if 
the resistance of the air nozzles or of the column of melting ore is 
increased. The centrifugal blower running at constant speed would 
not be able to drive the air through the mains at a pressure much 
higher than normal. If, therefore, no means for regulating the run- 
ning speed are available, some airangement must be supplied such as 
that furnished by Sautter, Harle and Cie, to the iron works at 
Chasse. Fig. 102 shows how coupling either in parallel or in series 
is combined with perfect balancing of the impellers. 

‘Although coupling the stages alternately either in parallel or in 
series permits us either to deliver a la1ge quantity at normal pressure 
or a reduced quantity at double that pressure, this mode of regula- 


TURBO-COMPRESSORS 125 


tion is unsuited to a great many cases. For instance, in the majority 
of chemical processes, the delivery of constant volumes of gas is 
necessary for the economical working of the process and the uniform 
quality of the product. One of the best known of these processes 
is the melting of pig iron in the cupola. The problem of delivering 


Fic. 102.—Arrangements for coupling turbo blowers. 


the constant volume against varying head can be solved only by 
using varying speeds, but under this condition it can be solved by 
the turbo-blower with almost unexcelled exactness. The means by 
which this is effected is an apparatus which has been called by its 
inventor, Prof. Rateau, the ‘Multiplicator.’ 

“Rateau Multiplicator—tFig. 103 shows a cross-section of the 
Multiplicator as applied in the well-known turbo-blower plant at 


‘Fic. 103.—Rateau multiplicator. 


Rothe Erde. The principle is the same as that of the Venturi water 
meter; that is, by tapering a pipe the velocity with which the gas 
passes through it is increased as the diameter is narrowed. As 
no addition or subtraction of energy is made during the passage of 
gas through the pipe, any increase of velocity must be accompanied 
by a decrease of static pressure. Ifthe air be tapped from the narrow- 
est and widest points of the tapered pipe, and the tapping pipes be 
led to opposite sides of a movable piston, it is clear that the difference 
in static pressure on the two sides of the piston will either move that 
piston or exert a certain effort on the piston-rod. 


126 AIR COMPRESSION AND TRANSMISSION 


“The greater the reduction of cross-section in the tapered pipe 
is made, the greater becomes the effect exerted through the piston- 
rod, and the greater also becomes the variation in that effort 
caused by any increase or reduction of velocity in the main; that is, 
of the quantity passing through this main per second. ‘Therefore, 
to obtain a very sensitive piece of apparatus powerful enough to 
move a governor gearing when the variation in the delivery is © 
but 1 or 2 per cent., it would be necessary to make such a great 
difference of cross-sectional area that the resistance of the mains 
would be considerably increased by the throttling effect of the taper. 


== x 

: ead £ 
= S 8 
Et Ny S 
C= ry ce 
3 

& 


Regulating _% 
Throttle Valve. 


7 
a) 


N , TZLILL ¢ 


Ngo oe 
Fic. 104.—Piston controlled by Fic. tos.—Connection between piston 
multiplicator. and regulators. 


“Here the ingenuity of Prof. Rateau’s method appears. He 
tapers the main but very slightly and inserts at the narrowest end 
a system of pipes as shown in Fig. 103. At the point a the static 
pressure is already reduced somewhat below that in the normal 
pressure main. At the point 6 the pressuie of the small quantity 
tapped off at a is again decreased. Finally, the static pressure at 
c is in turn much lower than at 6. The effect of this arrangement 
is so great that with a velocity of 60 ft. per second in the mains, 
the difference of static pressure between a and c was about 6 in. of 
mercury, while the loss occasioned by the whole installation was at 
the same moment not mote than 3/8 in. of water column. By put- 
ting the two ends of the cylinder A, Fig. 104, in connection with the 
narrowest and widest points of the tapered piping system, a consid- 


TURBO-COMPRESSORS | 127 


erable force can be exerted on the spring B. It can easily be cal- 
culated that a difference of 2 1/2 lb. in total pressure is generated 
by a variation of about 1 per cent. in the velocity of the air current 
passing through the mains; that is, when the quantity delivered 
by the blower varies by about 1 per cent., a force of about 2 1/2 
Ib. becomes available for moving the regulator. 

“Figure 105 gives an idea of the manner in which the gearing was 
arranged in a special case. The speed regulator and regulating 
throttle valve of a Parsons steam turbine were influenced simulta- 
neously. In like manner the regulating lever of any driving electric 
motor can be moved in exact proportion to the momentum of the 
air piston, as shown in Fig. 103. 

‘It is very interesting to see how this achievement enabled turbo- 
blowers of the Rateau system to create an entirely new field for them- 
selves. One of the licensees of Prof. Rateau, the machine-manu- 
facturing establishment of Kuehnle, Kopp & Kauschat Frankenthal, 
delivered some turbo-blowers for the Anilin and Soda Factory at 
Baden for the purpose of blowing air through the electric arc in the 
newly invented process of obtaining nitric acid directly from the at- 
mosphere. These turbo-blowels were to replace reciprocating blow- 
ers, which had caused great trouble and expense. It was necessary 
to connect them with a very large air-tank in order to produce reason- 
able steadiness of the air current, and when inspection of the recipro- 
cating blowers was necessary, it was impossible except by the skill 
of very experienced mechanics, and even then only with the greatest 
risk, to take one of the blowers out of service and at the same time 
start ancther without interfering with the continuous current of 
air. The turbine-blower not only gave an absolutely continuous 
air current, but proved so safe in operation that no change of blowers 
was needed during the entire process, which generally lasts uninter- 
ruptedly for several months. 

“Mixing Blower.—This extraordinary success of the turbo- 
blower impelled the Badische Anilin and Soda Fabiik to order 
fiom Kuehnle, Kopp & Kausch another kind of turbo-machine— 
that is, a mixing blower, which is shown in Fig. 106. Two different 
gases ale drawn by different sets of impellers, keyed on the same shaft, 
and are delivered to two different delivery pipes. This alrangement 
has the advantage that owing to the compulsory equality of speed 
of both impeller groups, the relation between the quantities of the 
gases continuously delivered is absolutely the same (that is, it is 
proportionate to the cross-sections of the impellers) providing the 
resistance remains equal in both delivery pipes. As this last con- 
dition cannot be kept uniform during the chemical process, auxili- 
ary throttle valves are inserted in the delivery pipes and worked 
by two Multiplicators. These latter, after careful adjustment, 


128 AIR COMPRESSION AND TRANSMISSION 

insure the maintenance of absolutely constant mixing rates between 
the two deliveries under all conditions. The makers were required 
to guarantee that the mixing ratio should be kept constant within 
a margin of 1 per cent., and their machines wete so perfectly designed 


iM 
an 


Ss 


Q 
Ly 
LY 
N 
K) 
k) 
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N) 
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N 


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5 cS vw ZZZZ csSh yy oavenh 
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Fic. 107.—Rateau turbo-compressor, 140,000 cu. ft. per hour. 


that the Badische Anilin and Soda Fabrik at once began to develop 
new processes for the electric synthesis of gases, which were made 


possible only by the new mixing turbo-blower.” 
The cross-section of a Rateau Turbo-compressor of 140,000 cu. 


ft. per hour running at 4,600 r.p.m. is shown as Fig. 107. 


CHAPTER XII 
HYDRAULIC COMPRESSION OF AIR 


The method of compressing air by means of falling water, without 
the use of any other moving part whatever, forms one of the most 
interesting topics in the subject of air compression. 

The large installations in northern Michigan, together with the 
large compressors of the same type in British Columbia, Quebec 
and Connecticut, give some idea of the extent to which this very 
simple method of utilizing the energy of falling water is being ap- 
plied. All of these installations have been completed within very 


Fic. 108.—The trompe. Fic. tog.—Frizell’s hydraulic compressor. 


SNe 


recent years and their success gives promise of many more such 
plants being planned. 

Trompe.—One of the oldest forms of compressing air is by means 
of a trompe or water bellows, a device of historic interest, in which 
water was lead from a higher to a lower level through a pipe or bam- 
boo pole with openings in the side through which air entered and 
mingled with the descending water and was later trapped from it, 
as shown in Fig. 108, for use in forges. 

A great many impovements have been made on this aan ap- 
paratus and quite distinct types developed from it. 

Frizell’s Compressor.—One of these is shown in Fig. 109, the 
invention of J. P. Frizell of Boston, Massachusetts. This device 
utilizes an inverted syphon having a horizontal passage C between 

9 129 


130 AIR COMPRESSION AND TRANSMISSION 


the two legs, Band. A stream of water is led into the upper end 
of the longer leg B and at the top of the horizontal passage C’ be- 
tween the two legs of the syphon, an enlarged chamber, D, is con- 
structed in which the air separates from the water. The water 
freed from the air passes up the shorter leg, F, of the syphon to the 
tail race. The pressure of air accumulating in the chamber is 
determined by the height of water in the shorter leg. 

This application of carrying upward the water after the air is 
separated from it seems to have been original with Mr. Frizell, and 
in this respect his device differs from the old trompe. 

Mr. Frizell made two working models of this type of apparatus. 
In the first one, the legs of the syphon were 3 in. in diameter, 
the head of water being 25 in. and an efficiency of 26 1/2 per cent. 
was obtained. A larger apparatus was then constructed at the 
Falls of St. Anthony on the Mississippi River a few miles above 
St. Paul; the longer leg of the syphon in this plant was 15 X30 in. 
and the shorter leg of the syphon 24X48 in. in section; the height 
of water above the air chamber was 29 ft. The head in feet varied 
from 0.98 to 5.02; the first head being just sufficient to cause a flow 
through the pipes. With the working head changed from 2.54 ft. 
to 5.02 ft., the efficiency varied from 40.4 per-cent. to 50.7 per cent., 
the quantity of water in these cases varying from 5.92 to 11.89 cu. 
ft. per second. 

Mr. Frizell estimates from the experiments he has made that with 
a shaft ro ft. in diameter, a depth of 120 ft. and a fall of 15 ft., the 
efficiency would be 76 per cent.; and with a head of 30 ft. and a fall 
of 230 ft. the efficiency would be 81 per cent. 

Mr. Frizell’s first experiments involved a large outlay in cost of 
plant and were not entirely satisfactory; but where there is a mod- 
erate water fall and plenty of water, this is no doubt a very simple 
method of compressing air. 

This system is applicable to either high or low falls and although 
no installations of this type of air compressors were made until a 
number of years after Mr. Frizell’s patents were obtained, the fact 
that he is the pioneer in this line entitles him to a great deal of 
credit. | 

The following explanation of this system is taken from The 
Railway and Engineering Review, Sept. 17, 1898. 

“The general principles underlying this method of compression is 
familiar to most in one form or another. For instance, it is well 
known how readily water breaks into foam, which is due to its being 


HYDRAULIC COMPRESSION OF AIR 131 


impregnated with airin minute bubbles. Since bubbles rise in water 
at a velocity depending on the size of the bubble, it is obvious that 
air drawn into a current of water moving downward with a velocity 
in excess of that at which the bubbles rise will be carried down and 
subjected to a pressure corresponding to the depth attained, and 
moreover the compression will take place isothermally, a process 
which is not accomplished by any method of piston compression. 
If the direction of the water be then altered to a horizontal one, the 
air will rise in a few seconds to the top of the passage and accumulate 
in a suitable chamber under the desired pressure. The length of 
the horizontal tunnel will be controlled by the necessity of placing 
the entrance to the air chamber far enough from the descending 
branch to admit of the complete escape of the air bubbles. 

‘‘A method of introducing air into the descending column of water 
is to surround the shaft with a bulkhead of masonry, over which 


Fic. t10.—Syphon bulkhead. 


the water is led in a covered channel, the bottom of which rises a 
little above the highest level of the water. This forms a syphon 
as shown in Fig. r1o. 

‘“‘At the point A the pressure within the syphon is less than that 
of the external air, and the latter will flow in through any opening. 
This is evident, because the flow of water depends upon the syphon 
principle. This space A extends around the masonry bulkhead and 
is in communication with the atmosphere. It is also connected with 
a pump for the purpose of removing any water that may collect in it, 
the amount of air being regulated by opening or closing holes in 
chamber A.” 

Baloche and Krahnass Compressor.—Another device, shown in 
Fig. rrr, differs somewhat from that of Mr. Frizell. It was invented 
by A. Baloche and A. Krahnass in 1885 and consisted of a syphon, 
B, carrying water from an upper to a lower reservoir, the lower end 
of the syphon being projected through an inverted vessel, R, placed 
nearly at the bottom of the second reservoir. Just beyond the bend 


132 AIR COMPRESSION AND TRANSMISSION 


in the syphon and in line with the axis of its longer leg, an air pipe, 
T, projected into the descending leg of the syphon. This entrained 
the air with the descending column and carried it down into the in- 
verted chamber, R, from which the air escaped at the top, while the 
water passed out from the bottom into the lower reservoir. This 
apparatus produced pressure on the air in the top of the inverted 
chamber due to the height of the water column upon it. 

Arthur Compressor.—Another device, shown in Fig. 112, patented 
by Thomas Arthur in 1888, differs from the last in having a stream 


Fic. 111.—Baloche and Krah- Fic. 112.—Arthur’s hydraulic com- 
nass’s hydraulic compressor. pressor. 


of water led directly into the top of the vertical pipe A. Inserted 
into the mouth of this pipe is a double cylindrical cone, C, forming 
an annular air passage between it and the walls of the pipe A. Owing 
to the increase in the velocity of the water passing through the nar- 
row throat of the double cone, air is inhaled through the pipe D, 
through the annular space mentioned and through perforations in 
the lower cone and is entrained with the falling water. 

Through the down-flow pipe A rises a vertical delivery pipe, Z, 
for the compressed air, having its lower end, H, enlarged and open at 
the bottom. Projecting upward into this enlarged air-delivery pipe 
is a water escape pipe, /, through which the water passes after 
parting with the air. The escape pipe is in the form of an inverted 
syphon and maintains on the air in the delivery pipe Z a pressure due 
to the elevation of the water at the discharge point above the air line 
in the large end of the delivery pipe. 


HYDRAULIC COMPRESSION OF AIR 133 


Taylor Compressor.—The hydraulic compressor system of Mr. 
Taylor is shown by Fig. 113. The large recent installations referred 
to are principally based upon his patents. 

With Taylor’s system a series of small air pipes placed vertically 
in the upper end of the falling column of water introduce the air into 
the water. The compressed air and water are separated at the 
bottom of the shaft. 

Mr. Taylor seems to have been the first to introduce the plan 
of dividing the air inlets into a great number of small openings 
evenly distributed over the area of the water inlet. 


Fic. 113.—Taylor’s hydraulic compressor. 


In the figure shown, these air tubes are represented at C, all 
terminating at the conical entrance B to the down-flow pipe H. The 
water supply is furnished to this down-flow pipe through a flume D. 
As the water falls it draws air through the small tubes, carrying it 
down to the separating tank G, where it is liberated at a pressure 
depending on the weight of the water in the vertical pipe Z. 

The compressed air is then conducted through the pipe K to the 
place to be used. The distance from M to the tail race L represents 
the height or fall of water that is available. 

Taylor at first seems to have attempted to utilize centrifugal 
action in causing the separation of the air and water in the large 
chamber at the bottom, but afterward abandoned this scheme and 
‘used instead forms of deflector plates in combination with a gradu- 
ally enlarging section of the lower end of the down-flow column in 


134 AIR COMPRESSION AND TRANSMISSION 


order to decrease the velocity of the air and water and cause the 
water to part more readily from the air. 

The position of the hopper or frame carrying the air inlet tubes 
regulates the amount of water that is admitted to the vertical pipe. 
The quantity of air regulates itself and is neither more nor less than 
the given quantity of water can carry. If the descending column is 
so loaded with air that it does not preponderate sufficiently over the 
ascending column, the water in the former will rise, the commotion 
will diminish and less water will enter. In the contrary case the 
water falls, commotion increases and more air is taken in. ; 

Taylor Compressor at Magog, Quebec.—The first one of these com- 
pressors on the Taylor principle was installed at Magog, Quebec, to 
furnish power for the print works of the Dominion Cotton Mills 
Company. The head of water is 22 ft., the down-flow pipe is 44 in. 
in diameter and extends downward through a vertical shaft ro ft. 
in section and 128 ft. deep. At the bottom of the shaft the com- 
pressor pipe enters a large tank 17 ft. in diameter and tro ft. high, 
which is known as the air chamber and separator. 

A series of very careful experiments have been conducted at the 
Magog plant by Professor Kennedy and others; and it has been 
demonstrated that with a head of 19 1/2 ft. of water using 4,292 cu. 
ft. of water per minute, the equivalent of 1,148 cu. ft. of free air per 
minute was recovered at a pressure of 53.3 lb. showing that of a 
gross horse-power of 158.1, 117.7 h.p. of effective work was used in 
compressing air, giving an efficiency of 71 per cent. which is very 
satisfactory. 

_This compressed air was then used in an old Corliss engine, with- 
out changing the valve gear in any way from what it was when 
adjusted for steam, and 81 h.p. was recovered, showing a total of 
work recovered from the falling water of 51.2 per cent. When the 
compressed air was heated to 276° before being used in the engine, 
11m h.p. was recovered. The heating required 115 lb. of coke 
per hour, equal to about 23 h.p. The efficiency, therefore, including 
the falling water and the fuel consumed, was 61 1/2 per cent. It 
has been calculated from other experiments that if the compressed 
air had been heated to 300° the total efficiency secured would have 
been 87 1/2 per cent. 

When it is considered that a good water turbine will give an eff- 
ciency of 85 per cent. and that part of the power developed in , 
the turbine will be lost through transmission before the power is 
available, it is evident that this system is a very efficient method 


HVDRAULIC COMPRESSION OF AIR 135 


of generating and transmitting power. For if the efficiency of the 
turbine is 85 per cent. and that of the system that is used for con- 
verting the power in the turbine into a more.available form 80 per 
cent., the total efficiency of the system will be 0.80X0.85 or 68 
per cent. 
This shows the immense importance of this device. Its field of 
usefulness is certainly a large one. 

Taylor Compressor at Ainsworth, B. C.—Figure 114 illustrates 
a sketch of the upper part of the Taylor Hydraulic compressing 
plant at Ainsworth, B. C., which is 


< A é 6 q ——oF 
quite unique in that it did not re- ) AT 
quire the sinking of a very deep B i B 
shaft. Theapparatusis constructed aAadianng 


against the vertical wall of the 
canyon in the rugged mountain dis- 
trict in which it was built. The 
plant is located on Coffee Creek to 
the south of Ainsworth and about 
2 1/2 miles from the principal mines 
to which it supplies compressed air. 
‘The creek has a flow varying from 
2,500 cu. ft. per minute to several 
thousand, and the flume used is 
stave barrel construction, round 
steel bands being bolted around it 
every at... Phe wtilume? is 17250711, 
in length, 5 ft. in diameter in the 
clear, the available head at the 
compressor being 107 1/2 ft. The 
water at the compressor tower is received in a wooden tank 12 ft. 
in diameter and 20 ft. in height. A down-flow pipe passes from 
the water level through the bottom of this tank down perpendic- 
ularly and at the creek level a shaft extends to a depth of 210 ft., 
making a total vertical height to the shaft of over 300 ft. 

This down-flow pipe, which is 2 ft. 9 in. in diameter outside, is 
also of stave construction throughout, the bands being placed 
from 6 in. to 3 ft. apart, depending on the pressure to which a 
particular section is subjected. 

This terminates in a great bell-shaped chamber at the bottom 
of the shaft 17 ft. in diameter and 17 ft. high, the bottom of this 
chamber being about 2 ft. above the bottom of the shaft, thus allow- 


Fic. 114.—Taylor’s compressor at 
Ainsworth, B. C. 


136 AIR COMPRESSION AND TRANSMISSION 


ing the water to pass out and up the shaft to the tail race. A deep 
circular groove was dug in the bottom of the shaft to aid in separating 
the air from the water. 

As the distance from the water level of the air chamber to the 
tail race is about 200 ft., the pressure on the air is about 87 lb. 
per square inch. 

The air is conducted from-the compressor through a g-in. pipe 
which supplies compressed air through several branches to over 
I5 mining properties. The total length of pipe is over 2 miles. 
A pipe reaches from the surface of the creek level—that is, the tail 
race—to the dividing line between the air and water of the large 
chamber at the bottom; so that if more air is being compressed 
than is being used, the water line in this chamber will be lowered 
and the surplus air escape, while if the pressure of air falls, the escape 
pipe will be closed. 

The actual effective head of water in the apparatus is 107 1/2 ft. 
and if a turbine had been used, with an efficiency of 75 per cent. 
the available horse-power generated would amount to 620. 

This installation has cost in the neighborhood of $60,000, including 
incorporation, water-power, development and pipe line. Of this 
investment, $20,000 will cover the pipe-line cost, $10,000 the water- 
power improvements, and $30,000 the compressor cost. This 
last item was unusually high because. of the extremely hard founda- 
tion through which the shaft was sunk. 

Taylor Compressor at Victoria Mine, Mich.—In 1906 a large 
plant of this type was installed at Victoria Copper Mine near Rock- 
land, Ontonagon County, Michigan, which consisted of three com- 
pressing units with a total capacity of from 34,000 to 36,000 cu. ft. 
of free air per minute. A series of tests made on a single intake 
head by Prof. F. W. Sperr, gave the following results: 


TABLE X.—AIR MEASUREMENTS 


: , Absolute pressures 
Velocity | Cubic feet |_ 
Square feet 

feet per per 


area 3 Free air, | Compressed 
second minute : 
pounds | air, pounds 


Horse-power 


4 44.09 10,580 14 128 1,430 
49.74 11,930 14 128 T,627 
4 38.50 9,238 14 128 1,248 


aN 


HYDRAULIC COMPRESSION OF AIR 137 
TABLE XI.—WATER MEASUREMENTS 
Mune Velocity | Cubic feet Bibcen cr: 
feet per per Head, feet | Horse-power 
area : per cent. 
second minute 
71.75 3-933 13,057 79.5 1,741 82.17 
67.03 3.684 14,820 70.0 1,961 82.27 
7210 2.936 12,710 70.6 1,700 7350 


Phenomena of Hydraulic Air Compression.—There are several 
phenomena in connection with this method of compressing air that 
at first thought seem paradoxial. 

In compressing air by hydraulic means, the air becomes drier 
during the compression, but no matter what may be its initial 
condition as to humidity at the end of compression it will, in all 
probability, be saturated with moisture. 

Air almost always contains moisture. Its capacity for moisture 
is determined by the combined conditions of pressure and tempera- 
ture to which it is at the time subjected. 

Changes, either of pressure or of temperature, immediately change 
the capacity of air for water, and if the free air is saturated with 
moisture its capacity for water will be reduced whenever the pres- 
sure is increased or the temperature decreased, and in consequence 
water will be precipitated. 

When air is compressed by hydraulic means, isothermal com- 
pression is secured and, generally speaking, at uniform temperature 
a given volume of air implies a capacity for a certain weight of 
water whether the air is at a pressure of one or one hundred at- 
mospheres, but if the air is compressed through a range from 1 to 
too atmospheres, its volume will be reduced, if the compression is 
isothermal, to 1/100 the original volume, and in consequence 
99/toc of the weight of moisture it originally held will be precipi- 
tated. The air is still saturated, but the total weight of water held 
in suspension has been reduced. That is, this method of com- 
pression has reduced the weight of moisture present in the air and 
hence dried it, but at the end of compression the air is saturated 
with moisture. 

Another interesting phenomenon in connection with this type of 
compressor has recently been brought to light. It has been found 
that air compressed by this method contains less oxygen than free 


138 AIR COMPRESSION AND TRANSMISSION 


air of the atmosphere and in consequence its use in mines is not as 
beneficial as air from other types of compressors. 

It will be observed that, with this construction, the material used 
for the down-flow pipe need only be of sufficient strength to carry 
the weight of water and pressure generated in the working head of 
the water-power, as once it reaches the tail race level the internal 
pressure is gradually neutralized from that point down by the pres- 
sure in the return water surrounding the down-flow pipe; so that any 
pressure almost may be reached without increasing the strength of 
the down-flow pipe. The material for the down-flow pipe may be 
iron or wood hooped with iron, and the shaft may be constructed 
of inexpensive timber as it is preserved by being constantly in the 
water. 

By this method, low falls, otherwise useless, are made available 
and the same pressures can be obtained as from high falls, the horse- 
power being determined by the diameter of the down-flow pipe, and 
the height and volume of water in fall, while the pressure depends 
solely upon the depth of the well or shaft; therefore, any desired 
pressure can be obtained. 

Briefly stated, the air is compressed by the direct pressure 
of falling water without the aid of any moving mechanism and prac- 
tically without expense for maintenance or attendance after instal- 
lation. 

By this system any fall of water varying in working head may be 
utilized, any pressure required can be produced and uniformly 
maintained up to the capacity of the water-power, delivering the 
compressed air at the temperature of the water. 

This drying of the air and the fact that practically isothermal 
compression is secured, form the great advantages of this system 
of air compression. ‘The initial cost need not be excessive, and as 
the cost of attendance is slight, for certain purposes the method is 
ideal. Its field of operation is quite broad, as a high fall of water 
is not essential, for any desired pressure can be obtained from any 
fall, the capacity being determined by the power available in the 
water fall. 

Losses of Hydraulic Compression.—The losses inherent in this 
method of compression are: (1) The head expended in impregnating 
the water with air. This usually amounts to about 1 ft. 

(2) A loss which may be called the slip due to the velocity with 
which the bubbles tend to rise. It is obvious that the rise of bubbles 
during the decent of the water is a lost motion which deducts from 


HYDRAULIC COMPRESSION OF AIR 139 


the efficiency of the system and in addition there is a head con- 
sumed in friction. 

(3) A loss due to the increasing solution of the air in the water 
with the increasing pressure as the water and air descend. 

This air does not separate from the water in the lower chamber but 
is eliminated in the ascending shaft in the same order that it is 
dissolved in the descending shaft. The escaping air in the ascending 
shaft aids the movement of the water and this partly balances the 
loss in the descending column. 

There are on the market to-day small hydraulic air compressors 
for furnishing compressed air in small quantities for dental and other 
purposes. They can be operated by water pressure from any water- 
works supply and on this account are particularly adapted for such 
purposes. 


CHAPTER XIII 
EFFECT OF ALTITUDE AND COMPRESSOR TESTS! 


As the density of the atmosphere decreases with the altitude, a 
compressor located at a high altitude will take in a smaller weight 
of air at each stroke, that is, if the compressor is located at a high 
altitude, the air is taken in at a lower pressure and in consequence 
the early part of the compression stroke is occupied in compressing 
the air from this lower density up to a standard atmospheric pressure 
at the sea-level. The reduction of pressure at the inlet would, of 
course, affect the power expended in compressing the air, but the 
decrease in power required does not vary in the same ratio as the 
decrease in capacity. For this reason compressors to be used at 
high altitudes should have the steam and air cylinders properly . 
proportioned to meet the varying conditions at different levels. 

Effect of Altitude on Capacity.—This matter is of special im- 
portance in connection with mining operations, because of the large 
number of mines situated in elevated mountain regions. The rated 
capacities of compressors, in cubic feet of air, as given in the makers’ 
catalogues, are for work at normal atmospheric pressure, and due 
allowance must be made for decreased output at elevations above 
sea-level. This reduction in output, which is usually also tabulated 
in handbooks and catalogues, should receive due consideration in 
order to avoid serious errors. For example, the volume of 
compressed air delivered at 60-lb. pressure, at 10,000 ft. elevation is 
only 72.7 per cent. of the volume delivered at the same pressure by 
the same compressor at sea-level. In other words, a compressor 
which at sea-level will supply power for 1o rock-drills, will at an 
elevation of 10,000 ft. furnish air for only 7 drills. 

Effect of Altitude on Power.—The foregoing statement relates 
only to the volumetric capacity of the compressor. It must be re- 
membered that the heat of compression increases with the ratio of the 
final absolute pressure to the initial absolute pressure. As this ratio 
increases with the altitude, more heat will be generated by compres- 
sion to a given pressure at high altitudes than at sea-level. This 
additional heat temporarily increases the pressure of the air in the 
cylinder while under compression, and more power is therefore re- 


1Peele, Compressed Air Plant. 
140 


HHPECIOCOFr ALTITUDE AND. COMPRESSOR TESTS ~ 141 


quired to compress and deliver a given quantity of air. The cor- 
responding loss of work, due to the subsequent cooling of the air in 
receiver and piping, also increases with the altitude. 

Relation between Altitude and Volume.—Contrary to a common 
impression, the volume of air delivered by a given compressor does 
not bear a constant ratio to the barometric pressure, but at different 
altitudes this volume decreases slower than the barometric pressure. 
This relation may be shown as follows: Two ideal indicator cards are 
represented in Fig. 115, one of a compressor working at sea-level with 


= 
_ 
-—- 


=<- 
“= 
oa” 


4 
Ce 


P. 
P. 


—----—— 


eth eet be as ee 

1 

' 

| 

{ 

t 

t 

1 

f 

< 
4 re 

, 

| 

| 

1 


Fic. 115.—Effect of altitude. 


an initial pressure Py, the other at an altitude with a lower initial 
pressure P2. The initial volume V and the final gage pressure P 
are the same for both compressors, P3 and P, being the respective 
final absolute pressures. V1 and Ve are the final volumes, corre- 
sponding to the dotted isothermal curves, these volumes being 
taken as the basis because they are those to which the compressed 
air will eventually shrink on losing the heat of compression. From 
the theory of air compression, 


VAeP 
VP\—= Vik; or Wire, (1) 
Vee: 
and VPo=VoPa, or V7, (2) 


But since P3=P,+P, and Ps=P2+P, equations (1) and (2) may 
be written: 


ary Cae if je (3) 


and is ena. i (4) 


142 AIR COMPRESSION AND TRANSMISSION 


Dividing equation (3) by equation (4) 


li 
i+ 
Vogue 1 F Bae i : ge 
ae po Vie Vee ioe pease (5) 
ari 


This gives an expression for the ratio between pressure and volume 
at sea-level and for any altitude above sea-level, of which the corre- 
sponding barometric pressure isP2. Thus, let Pps=10lb.,P=golb., 
and V;=0.1404 cu. ft. By substituting these quantities in equa- 
tion (5), V2 is found to be 0.0999, or approximately o.1 cu. ft. 

In Table XII, columns 4 and 5, are given the relative volumetric © 
outputs, at gage pressures of 70 and go lb. of a compressor working 
at different altitudes, the figures being percentages of the normal 
output at sea-level. These percentages have been derived by 
Mr. F. A. Halsey from equation (5), a constant loss of initial pressure 
of c.75 lb. being assumed to allow for the resistance presented by 
the inlet valves, to which reference has been made in another 
chapter. That is, for practical purposes the sea-level atmospheric 
pressure is taken as 14, instead of 14.7 lb. The other columns 
show the mean effective pressures and indicated horse-powers, 
corresponding to different altitudes, up to 15,000 ft., which will be 
found convenient for reference. It should be noted from the figures 
in columns 4 and 5, which are for the ordinary range of pressure 
employed in mining, that, though there is a difference of 20 lb. 
between the two gage pressures, yet the outputs at different altitudes 
vary only by a few thousandths and may often be neglected.! 
Wide differences, however, occur in the columns of mean effective 
pressures and horse-powers. 

Owing to the increase of piston displacement per indicated 
horse-power, as shown in columns 8 and 9 of the table, some 
builders make the air cylinders of compressors for mountain work 
of larger diameter for the same size of steam cylinder than those 
for sea-level service. As against the losses of the air end of the 
compressor at high altitudes, there is some gain in mean effective 
pressure of the steam cylinders, because the exhaust takes place 
against lower atmospheric pressure. The same is true in part 
of the air exhaust of machines using the compressed air. But the 

1 Attention may be called to the fact that for this reason, in compressor- 


builders’ catalogues, no account is taken of the gage pressures in tables of 
compressor capacities at altitudes. 


EFFECT OF ALTITUDE AND COMPRESSOR TESTS 148 


resultant of these gains is small and cannot be given much weight 
in offsetting the losses. 


TABLE XII 
. ‘ f , 
Barometric poke eee Cubic feet 
Pressure pistons compressed 
.7 Relative out-| M.E.P. for placement : ene 
: ce. oi air per indicated 
put for gage gage pres- per indicated 
horse-power 
Altitude, pressure sure horse-power foreraze 
feet Pounds for gage 
Taches ve préséitre pressure 
mer- 
square 
cury ach 
: 7.OnlDeiLOOm Delle Onl Den OO; bs 1770: 1bn00) Lbs 1.70 1b. | 90 1b. 
I 2 3 4 5 6 i. 8 9 Io II 
fo) 30.00 17 5S OOO LE NOOOU Sale So Neal OOS 5,00 et. LAAN On sor 
I,000 28.88 Te? ONO TOO 7aLOR OOO IES 2e Ollme 37> Olle yn O3 ale O,. 00m i. k23 nl cOm7 a7 
2,000 27.80 TA OVMMOMOS Su LOTOS ses 2a. MesOnOM) 7S Os2On) ele Toss OL773 
3,000 26.76 U3 eLOe | OFOOA LO ROOOH EST a Sle 3053) I) Feo 71, Oc ZT Hl T.O84) 190.7750 
4,000) 25.76 T22O7MOnS (nO eCOON, GEeO leas eOmle7 43004) Oo4 30 2. O05 a) 0.740 
5,000 24.79 P22 OMNLOMGAG HOSS Ones On5 ues 5 Omiya 5 Loos 55 aie BOAO) Pron 733 
6,000 23.86 Tey Ga ROMO Css LOMmOO Olle OOM s Aisin OS) al. Or. Oil ate O26) \kOe 720 
7,000 22.97 Dees On One Sun Ole OON ee Oud aime 37 lee 1S Onl O47 ON ie ON Le sOn 70'S 
8,000 PAP) Su TOMS 7 OMT SOULOM 7S Lm2o Oma sa ale 704 O02 OL OOd i) On005 
9,000 PA PAO, LOMAOM MOR 72 TMPOM 7232 Seow S2n Silo 5OOMs 7: OONl O1O76 ale 0.083 
I0,000 20.49 TORO 7a OMT OSL OnOOO! E276 5. lest Sil (Oo 224 071.204) OLOSOMI T0070 
II,000 Lon 72 On OmmFOF OSOML ORO Mae 2 7 Amie Gl so 3O ls 7134 aOLOAZE I O2058 
I2,000 18.98 Onset LO sO5010 50473 |620,.0)1 30.0.9) 8554 107240, 1) 0.025 5) 0.046 
13,000 Qua] SAOSMLOVOS2IOmO 232003 sO MOMS. 7k wn OAM OOO Ga 1041035 
I4,000 D750 865 04008) 02.000)) 25q8 e204 oe S58) 7.800) On8Org 02624 
15,000 16.93 SES 2 MOMS osNLOM S57 OlmMe Se sale COOL OOM: 7 (OO O.o 75 moe OLS 


The relation between compressor output and barometric pressure 
may be expressed simply in another way. Take the case of two 
compressors of the same size, one operating under an atmospheric 
pressure of, say, 14 Ib. and the othe at 10 lb. (corresponding approxi- 
mately to an altitude of 10,000 ft.). If the first compressor is 
producing 6 compressions, the final absolute pressure will be 146 
= 84 lb. or about 70 lb. gage pressure. To produce the same gage 
pressure, the other compressor must work to an absolute pressure 
of 7o+10=80 lb., the number of compressions corresponding to 


ep ibe 2 
which is ert From each cubic foot of free air the compressor 


will produce 1/6 of a cu. ft. of compressed air, and the second 
complessor, 1/8 cu. ft. Hence, the ratio of the respective outputs 
of the two compressors will be 1/8+1/6=3/4 or 0.750. As com- 


144 AIR COMPRESSION AND TRANSMISSION 


pared with this, the ratio of the respective barometric pressures 


5 10 
iS ade 


COMPRESSOR TESTS 


To indicate the observations required to secure the data for 
the complete test of a compressor, together with the deductions 
from the observed data, the following record of the test of a com- 
pound, two-stage Nordberg compressor, at the mines of the Tennes- 
see Copper Co., will be found useful.’ It will be noted that items 
28, 29 and 32 to 35 are necessary in this case, because the boiler 
plant supplied steam for the hoisting engine and an independent 
condenser, as well as for the compressor. Though the hoist was 
not running, steam was passing continuously to the jackets of the 
cylinders. The same conditions would often be met in other tests. 
The boiler-feed water was taken from a wooden tank, and during 
the run this water was supplied from two barrels on scales set 
temporarily over the tank. The water of condensation from steam 
jackets and reheater was drawn off continuously and also weighed. 
The calorimeter tests were made with a Peabody throttling calori- 
meter. Eight sets of indicator cards were taken during the 8-hour 
test, at hourly intervals. 


ITEMS OF COMPRESSOR TEST 


Altitude, 1,800 feet 


1. Date of test, February 16, 1902. 


2. Dutation of.test. hours<,6) a 8 
3. Diameter of high-pressure Sank (aindee 
(steam jacketed); inches im wen ee eee 14 
4. Diameter of low-pressure steam cylinder 
(steam, jacketed), inches.) v)j.s.42 4am. 28 
5. Diameter of low-pressure air cylinder, inches... 24 1/2 
6. Diameter of high-pressure air cylinder, inches. 15 3/8 
7estroke ol allpistons, Inehesa 4. ase eae ee 42 
8, Diameteriol piston-rods<inchés.. eee 201/16 
g. Revolutions of engine, average per minute.... go 
10. Piston speed perminute. ect. .n) al eee 630 
11. Steam-gage pressure, average, pounds........ 145.9 
12. Temperature of steam in bees average 
degrees Fahrenheit. . ee 364 


1 Abstracted from an article by I Parke aera Mines and Minerals, May, 
1905, P- 475- 


EFFECT OF ALTITUDE AND COMPRESSOR TESTS 


10 


rat 
14. 
15. 
16. 
17. 
THe) 
IQ. 
20. 


2 


22. 


Steam pressure in Spee Seeitee average 


POUDCSe eaten he ean 8 
Vacuum in Pordencer™ average inate BA aes oer 
Air pressure in intercooler, average pounds... . Dee 
Air pressure in receiver, average pounds...... 70. 
Temperature of air at intake, average degrees 

Fahrenheit. . foes 65. 
Temperature ie air sieving Noaenresetce eyline 

der, averave degroesahrenneit, .,...40.. 201" 
Temperature of air leaving intercooler, average 

deprecsalreinrenhiel ie sie teh tee ies 78. 
Temperature of air leaving high-pressure cylin- 

gderrvaverave cevreesm abrenneth: j) vo... ss , 240. 
Indicated horse-power in high-pressure steam 

CYMNGCE AAVCLACG at an eae, situs ne. oh 140. 
Indicated horse-power in low-pressure steam 

CyUnder Aa Vera ge meme ieN pe teen, an abel. oath. 153. 

. Indicated horse-power in both steam cylinders, 
PVCEAC Owe an ore eae 203. 
. Indicated horse. power in plows -pressure air ee 
ITC CUM VEL APC ARP ete et C00 vay Ci ddacs ae a 743" 
. Indicated horse-power in high-pressure air 
CVILGe Lercay CLAP ern city a euels awe da nets eh a Tage 
. Indicated horse-power in both air cylinders, 
VOU Omget, pcre Men Sole a ede eS eue LARP LS 278. 
. Feed-water weighed to boilers, pounds........ 43,343 


. Re-heater and jacket water from compressor, 


Were edt OUh Co Auer ee Peat ides Ade gee 4,081 


. Average temperature of re-heater and jacket 
water, degrees Fahrenheit....'...... 350. 
. Total heat in 1 lb. of steam ae 200d ve F., bheat 


LARUE ete Ea fers eR ly rele acct Sea at a cup 1,190 
. Total heat in 1 lb. of water at 356.7° F., heat 

units. ids 320. 
3 amvalcat edie tis re- rieiter anh Ayaas 

WaterepOUNdSr vate tear. Sai? Rae Le 
. Water weighed tron Rosideiseen orict in Renee 

engine acket, DOUNUS#. fe. ele ae 1,781. 
. Steam used to run condenser, pounds......... 4,320. 
. Total credits to feed-water, pounds.......... 7.228). 


. Total feed-water charged to engine, pounds... 36,115. 
. Moisture in steam shown by Peabody calor- 


imeter, per cent.. MR att ee or on Coes hae eh en iy 
. Credit for ribistaret in eecenn sounds ey crea & Aus 
. Total steam charged to engine, pounds....... 35,642. 


. Dry steam per hour charged to engine, pounds.. 4,455. 
. Steam consumption per indicated horse-power 
Dev MIOUL MPOUNGS an eal. saath al OM neckdye sf. 5, 


66 
63 


I2 


O02 


8I 


OO 


oOo 
oOo 
OO 


30 
OO 
OO 
0O 


19 


145 


146 AIR COMPRESSION AND TRANSMISSION 


42. Guaranteed steam consumption per indicated 
horse-power per hour, at 92 revolutions per 


Taintite pounds.) aa en. 14.00 
43. Excess of steam SORE TRD TORN per cued 

horse-power per hour over guarantee, pounds 1.19 
44. Theoretical delivery of free air per minute at 

oo revolutions; cubic teet.|. 2)... ke ee O59 oO 
45. Slip of air (percentage)... sok gates els) 
46. Actual slip of air per fnute. vcabic tenes As HAE A Gilet 
47. Actual delivery of free air per minute, average 

CUDICH ECL aay cre: etLs07 0077 


48. Theoretical terse -power Prenrede teh compress 

and deliver actual delivery of air at receiver 

pressure by adiabatic compression. . as 306.53 
49. Theoretical horse-power required to compress 

and deliver actual delivery of air at receiver 


pressure by isothermal compression. . ee 229.00 
50. Actual horse-power shown by air rican 

cards.. eee 278.81 
51. Actual eee -power howe bis steam i eevee 

Candseeeak teu 293515 
52. Actual Ree -power teonoumed iy fockon of 

Celie eee : 14.34 
53. Efficiency reo betw eens steam Pend air Rasatie 

ders, petscel tye ae ee OSes 
54. Efficiency ratio henween tear cae air veoiee 

ders guaranteed by builder, per cent. 87.00 


55. Efficiency of steam, or ratio of steam indicated 
horse-power to theoretical air indicated horse- 
power, isothermal compression, per cent.... Yow f 


One of the combined indicator cards, from which the averages in 
items 21 to 26 were calculated, is shown in the upper part of Fig. 116. 

A series of tests were made in 1909 by Richard L. Webb, consulting 
engineer, of Buffalo, N. Y., on a large number of compressors in a 
well-known Canadian mining district. In conducting these tests, Mr 
Webb had access to plants which have been in operation for a year 
or more under normal working conditions, and his results are of 
value not only to users of air compressors, but also to the manufac- 
turers. As a rule, the plants tested were in the care of competent 
machinists and in good running order, so that the results obtained 
may be taken as representing a fair average of current practice in 
the United States and Canada. The results of a few of these tests 
are given here to show the importance of determining the actual 
efficiency of air compressors when working under the conditions pre- 
vailing in most mines. 


EFFECT OF ALTITUDE AND COMPRESSOR TESTS 147 


Discharge Pressure 


x 
GA 
ZO" C163 % 


Scale 20 


Fic. 116.—Combined cards from two-stage compressors. Upper cards from 
Nordberg compressor. Lower cards from Ingersoll-Rand “Imperial Type 10” 
electrically driven compressor. Air cylinders 23’ and 14” X20”. 


ROU tONS Per NINE Ca ete eer MED eye eee re Bane. atr een) BOT 
Histon specd tect per minute c.f... ee ole we. wes es O23 23 
Discharge air pressure, pounds...................--.. 93 
Tntercodher pressure, DOUNdSs, o atau ustincy ts ok te es 24 
Volumetric efictency (from catd)og cata. iss... | 9563% 
Teeter ol low-Dressitercy Gehan ie tet he es Marg E32 
etl Peonich-pressure cylinder. 22 40 te. et) T2078 
BERL Ci teeter ee ae eae ee SASS a aden ASA. 8 
Free air delivered per minute, cubic feet (from card)... 1706 
HiMClenCY: COMP areas WitOyAGIAD pie a: oo sues Sis ee << O72 0 
Pecieney COMarER WIL ssOtUae we crus a aaa fares nue oo) OFM RVS 


Mode of Conducting the Tests.—The following plan was employed 
in each case. First, a boiler test was run for not less than two weeks, 
the coal being carefully weighed, the boiler feed-water measured, 


148 AIR COMPRESSION AND TRANSMISSION 


and the total revolutions of the compressor recorded by a revolution 
counter. From these data the cost per boiler horse-power and the 
average speed of the compressor were determined. Second, the com- 
pressor was operated at various speeds over its entire range. By 
means of a meter installed in the steam-pipe near the throttle, the 
total steam consumed, in pounds per hour, was measured. Indi- 
cator cards were taken on all cylinders, together with temperatures 
at the air inlet, intercooler, and discharge. To measure the actual 
volume of air delivered, a meter was placed in the discharge pipe 
outside of the receiver. A number of simultaneous readings on all 
instruments were taken at each speed. From these were calculated 
the total horse-power of the steam and air cylinders, the steam 
consumption, and the total piston displacement per minute. 

The air and steam meters were of the Dodge type, as modified 
by the General Electric Company, and were operated by their ex- 
pert sent for this purpose. The indicators were of the Roberts- 
Thompson and the American-Thompson make, which are well 
known and generally accepted as standard. Their springs were 
calibrated by a standard gage. 

Results of the Tests.—As was to be expected, the friction loss was 
found to be only a small item in the total. The other losses, which 
are frequently overlooked or disregarded, played a large part in 
cutting down the efficiency. The capacity of air compressors is 
usually rated according to the volume of the cylinders. On this 
basis, the mechanical efficiency only is given. For example, if the 
horse-power of the air cylinder is 100 h.p. and the horse-power of the 
steam cylinder 110, the efficiency of the compressor is rated as g1 
per cent. This rating disregards the losses due to adiabatic com- 
pression, heating of the cylinder and friction of the inlet and delivery 
valves. The tests show the friction loss of the engine itself to be 
usually not less than to per cent., and often considerably larger. 
Losses from the other causes mentioned were found to range from 
20 Per cents up: 

As Mr. Webb is not at liberty to disclose the identity of the partic- 
ular plants at which the tests were made, each test has been desig- 
nated by a number. 

Test of Plant Number One.—This consists of three 125 h.p. 
return tubular boilers (one being held in reserve), supplying steam 
for a cross-compound condensing air compressor of standard make. 
The steam cylinders have Meyer valve gear and are 16 in. and 28 in. 
diameter by 24 in. stroke. The two-stage air cylinders are 28 in. 


EFFECT OF ALTITUDE AND COMPRESSOR TESTS 149 


and 18 in. by 24 in. From a two weeks’ run the following results 
were obtained: 


Fas esas al ie ee el 


‘ 
= RIDES aaa 
Se ES 
| aR Ne . 
HEBRUSY 
Eide tases s 
ramen ' 
(aie iasltviesl ave i 
Rene ase : 
honwoeZes : 
Cet Aone 
ZZ caie win 
Wa : 
0 10 20 30 40 100 110 


R. p.M. 


Fic. 117.—Test results of compressor plant No. tr. 


Potaltcoalspurneds pounds s255as. easels en «2 2 204,300 
Total iced-water scubiCyeGtsm. fort en err ss oy. 37,459 
Total feed-water, pounds.............. . 2,335,508 
Average temperature of feed-water, degrees 


Hahrenhiel tage seman on wg ers eee Mach 131 
Average evaporation per pound coal con- 

SUTMCC MDOUMCS ae tue Sor s Thee, weoah te mnce. ak S72 
Average revolutions per minute............... 03 aE 
Indicated horse-power of steam end, corre- 

sponding; to 6321 t.p.1, af 2 5 yom sean - 161 
Corresponding indicated horse-power of air end. 123 
Average steam pressure, pounds.............. £15 
AN OLAGesVaCUUIh,, DOUDCS, 028, 74: ajacteere ss sock TO.5 
Averageyail pressure, pounds, 6:5 hte ness 96 
Average temperature of outside air, degrees 

PATI TORLELE Roe Patan RRO cases ocala: ohe/4 24 
Average air piston displacement at 70° F., 

CUDIC LECL ante ou eee Wet Cats hen iat pe poh wye.as Tia 


Average metered output corrected to 70° F., 
CU DEGHLOOL See mea ibe td direst beets Sat, eb wnat sl 758 


150 AIR COMPRESSION AND TRANSMISSION 


The average evapoiation of 8.72 lb. of water from 131° F. to an 
average steam pressure of 115 lb., is equivalent to 9.83 lb. of water 
evaporated from and at 212° F. per pound coal consumed. At 
the average compressor speed of 63.1 revolutions per minute, the 
metered output was equivalent to 758 cu. ft. of free air per minute, 
the piston displacement being 1,172 cu. ft. per minute. Fig. 117 
presents some of the principal data of the test run on this compressor. 


A -CostofCoal perf. HP perAnnum... 


Tanaris B- * and Labor- 


C-\ 9° "| Labor, int and Dep. 
D-Total Cost per Annum. 


Ace Ae 
Poe mol BOd BEE ees Ge ee ee 


IPAS 


WAS VAs 


ONO TB fea etc | cee 


ieee] SS esl | a} [SY Aas ee ae] 
BEES ibaa Boees 


oe Vee fens |e ea 
i enamel Ys Pa a Soe | 


ied a Ls Fe eS 
seh eel oe Feta i) Sl tla ay ete 
Fea og a i a ee 


r) 


Fic. 118.—Test results of compressor plant No. 1. 


To find the average operating results, the curves at 63 revolutions 
should be followed, at which the indicated horse-power of the steam 
cylinder was 160.8, and that of the air cylinder, 123, showing the > 
mechanical efficiency to have been 76.5 per cent. ‘The theoretical 
horse-power required to compress isothermally 1 cu. ft. of free air 
per minute to 96 lb. (the average pressure) isc.129. The theoretical 


HEP ECT OF ALTITUDE AND COMPRESSOR TESTS 151 


useful work done by the compressor is, therefore, 758 X0.129 or 
97.8, and the net total efficiency of the compressor is 97.8-+161 or 
60.8 per cent. 

The cost data were furnished by the owner and are based on one 
year’s operation. In Fig. 118 these costs are plotted, showing 
how the cost per steam horse-power, per year is affected by the 


$0 


i plea ep 
Nea SNCS Goooeeeeee 
EN Saale beets 
tas pam melanie taal 
eal Magee menienal ik 
ied eesti ri 
elie roa ees estan Poe | 
Sh ele FUG a 
ait Smo e les iaia 
ue SS CEA ee a 
on Se eeeeee res 
mises 
1a a 
Stl Recieve: 
ac ela RRA REM 
2 ail Mieisiehs elalelshlal 
Se lise: Remmi alsin 
SHER Ft 
eu ia Baer DEL 
Beaeld seroeiile lab ieee 
ia ies Melosh | 
abd Song Taam 
i Hloe a oP ST 
iia is ce Pe InP JN 
fafa a eae a ae IT 
esa Lo | 
R. pM. > 


Fic. 119.—Test results of compressor plant No. 1. 


average running speed of the compressor. The curve of Fig. 
119 shows the operating costs in another way. ‘These costs may 
be read in terms of 1,000 cu. ft. of free air compressed to too |b. 
or 1,000 cu. ft. of compressed air at too lb. gage pressure. 

Test of Plant Number Two.—The plant consisted of three 
150 h.p. return tubular boilers, supplying steam for a Corliss engine, 


152 AIR COMPRESSION AND TRANSMISSION 


the air compressor, and steam heating. To determine the boiler 
horse-power, a meter was placed on the steam-pipe to the compressor 
during the test run, so that only the portion of steam actually used 
by the compressor was charged to the same. The compressor was 
duplex, with Meyer valve gear, simple steam cylinders 14X22 in. 


real ats ee | | a pe 
of TT A a 
ce a A 
STA an ie 
a] ee | ea 
BB 02 oo 
wo EEA g 
A 5 0g 
TSE 4 caer: 
A A A 81s 
le 1 a lays relies 
oe in EE AAD, ail 2 
BURN a bales 
i MPV per 
x 5 x 
- 4 fe 70 e a 
HEME BAVOGGn ears 
BG MOO Anime. i.e 
RIGA SIRE, a 
HOPE He tat 
h/ 3 
IM WAU/ ApS Cle E ee). in 
Bar 
30 100 
20 0 


0 20 40 60 80 100 ~=3=—-:120 
R. p.M. 


Fic. 120.—Test results of compressor plant No. 2. 


and two-stage air end, 14 and 22 X22-in. stroke, rated by the manu- 
facturer at 1,050 cu. ft. of free air per minute at 105 revolutions. 
At this plant the test lasted over a month, with the following 
results: 


EFFECT OF ALTITUDE AND COMPRESSOR TESTS 153 


otal coal consumed apounds.. <n. a. ti. 05,2 e0rau 450,250 
LOCAL Leeds Water epOUunUs tea oo eyes suctod wt 2,496,000 
Average evaporation per pound coal consumed, 

POUNGS 5 ie eee eee atte take Sheets 5.46 
Average revolutions per minute............... 36 
Corresponding average Micatec meaner 

CFrOMIECUEVG) to, Warren ian, ee ee ead 53 


HHatHt+ee 5 daborn Gs Pacts in 
2 ee cate [kN lew RAS Ob Le | 
= SCENE EEE 
> NIC e ANG eee ee 
: ACO Ce ao ae 
E aaah e GRanee cnn dea 
> 
ena | pel 
a eilghel 
Serta iat) 
adit a bess Chad OE 
ia) Be Wa S2S0RR an 
eo Les Siete 
= LS De aes es 
: Cease 
5 aba Ca a el ae 
a a ex eesti ee Ves ees 
= 100 
a 


0 10 20 30 40 50 60 70 80 90 100. 
R. p. M. 


Fic. 121.—Test results of compressor plant No. 2. 


Hourly readings of the revolution counter were taken, showing 
an average speed of 36.05 revolutions. At this speed the steam 
consumption was 51 lb. per indicated horse-power hour, as measured 
at the throttle, the air meter showing a delivery of 275 cu. ft. of free 
air per minute. The total efficiency was 67 per cent. Taking the 
ordinary method of computing the mechanical efficiency only at the 
same speed, there would be 48 air h.p., divided by 54 steam h.p., 
giving an efficiency of 89 per cent. 


154 AIR COMPRESSION AND TRANSMISSION 


The coal consumption per indicated horse-power per year, as 
shown by the books of the company, amounted at the average speed 
to about 56 tons. Figs. 119, 120, 121, and 122 present details of the 
test on this plant, which was conducted similar to that on plant 
No. I. 


—152 
[eg a sale 
4 Ra Be KI de vali) 
ai 
PEELE EEE} 
VP) 
o 
i 
So 
S 
7 Se 
S 
El aN IS ICS 
7, SRR RA GR BRAS EAS maee 
Qo. 
= 
eh EERE EERE 3 
rd UE EMI ASE lL 
2 SSCS ese eee 
oO 
SII F i ae Sa ies eee 
We #2) 
ea aw Reo lalae 
ea ea BAG elaciealees alla 
5 [aa ae aa | elo 
seein 2 
oa eae BS WMEIDILI Le 
EE Se 


AY) 


Fic. 122.—Test results of compressor plant No. 2. 


Test of Plant Number Three.—This plant consisted of two 125 
h.p. return tubular boilers, supplying steam for a non-condensing 
cross-compound air compressor of standard make; steam cylinders 
18 and 35 X24 in., air cylinders 14 and 28X24 in. A two weeks’ run 
gave the following results: | 


Total-coalbutned, pounds j-\..) eee 221,190 
Total fced-water, cubic feetiay a. sue nas fae 7a 
Total feed-waters pounds .f20.7) eee 2,094,057 


Average temperature feed-water, degrees Fah- 
Tenhel tie. chaos cue OM Pe wih ee ar 154 


EFFECT OF ALTITUDE AND COMPRESSOR TESTS 155 


Average evaporation per pound coal con- 


SUMMER SOU Ge wager net yc ents ira ster at. wis 9.48 
Avetage boiler horse-power.. 0...’ s05 fag ees 208 
Average revolutions per minute: ..55 2.0. 06 cc. 66 
Average indicated horse-power of steam end, at 
OUT DAMM aloe Cury Gy maa ons teat ok avg es 210 
_ Average indicated horse-power of air end (from 
CULEVG) RANE eRe Ie, ore haan Naa T2005 
Ayeracersteam preseureayasue nadine 22s ee eos oon 97 
Average Gi Pressure sje. wet seta tae yas plas ee 07 
Average outside temperature, degrees Fahren- 
eT teat ete Cee emer nr AE cred SM cee Lig 3 . 23 
Average air piston displacement at normal 
SPecd cubic resteati goa Hea: af ects Wes Tey 
Metered output in cubic feet corrected to 70° F. 734 
4 CHES EE<a 
[SIE eas aes a a 
Pa CEES PAE Pe CS ae 
Ee] ee -eA--RER-- H+ VAAN head | 
Five MeMowmMGmmlei7icitee sk 
PERE Aga . 
pe Ie ea vac = 
Beet acgU lace 5 
0s el ES Sa FaRURSEs & 
oh [Bil SBE I Sala Laeadicl| alles 
eet (ea aT a a | 
ee iS A) gel ats ® 
BERS Zcn eee alee 
daa Es Ca is wa he 
ott} | A ea ore, patel | | bos As00° 
YN eh cet +} ° 
ara eee aa oe 
aaa CS ESP aU A 
PEAR be ke sae 7 
re Le | YA Pe Like ae fk CUE aE ie 
0 10 20 30 40 oat be ce 70 80 90 100 


Fic. 123.—Test results of compressor plant No. 3. 


The average evaporation of 9.48 lb. of water per pound of coal, 
from 154° F. to an average steam pressure of ¢7 lb., is equivalent 
to 10.4 lb. of water evaporated from and at 212° F. At the average 
speed of 66 revolutions, the displacement was 1,240 cu. ft. of free 
air per minute, while the metered output was 734 cu. ft., showing a 
net volumetric efficiency of 59 per cent. 


156 AIR COMPRESSION AND TRANSMISSION 


To determine the conditions in average operation, the curve at 66 
revolutions should be followed (Fig. 123), at which the indicated 
horse-power of the steam cylinders was 210, and that of the air 
cylinders, 128. This shows the efficiency to be 61 per cent., the 
friction loss being 81.5 h.p. or 39 per cent. of that delivered by steam 


Be GR 
se [Pa Ba Sa ce. saya arid Depreciation: 


+ ,Depn, Repairs & Supplies. 


\ 
<£)oE) Sas an aa 
af a a 
BPS EI apart 
Re en — 2 PARSE RRP | 
FE ete falc ec 
aE} ee 
BSE Vin OM EORGLE sae 
SA a CPL 

aia Nate ste [art aiat aaa a 
oT =aTaN Neal aa 
BUEN CERRORESE ROLES 
he al 
BERGEN ce Sec ce 
fa] Nes nT ISIS a ea a 
HN BRNSSEE 

Pi STASIS IOASS 
Mars Ee 

es naa ST 


0 10 20 30 40 50 60 70 80 90 = 100 
R. p.M 


per ILH.P per Hour 


S 


Dollars 


Nae 
ie 
NN 
ESS 


NI 
Ss 
EI 


Fic. 124.—Test results of compressor plant No. 3. 


end. This extremely high friction loss was due to the fact that the 
compressor shaft was out of line, and the plant could not be shut 
down long enough to rectify it. The details and results of this test 
are interesting in exhibiting the inefficiency that may be caused 
by a purely mechanical defect. (Figs. 123, 124, and 125). 

Test Number Four.—The results of a test on another plantare 
given in Fig. 126, the details of the boiler test and of the costs 
being omitted. In this case the compressor was of the tandem 
compound non-condensing type, with Corliss valve gear for the 


PEP ECTAOR ALPIRODE AND COMPRESSOR: TESTS Se157 


steam cylinders. The test shows that, at a low speed, the steam 
consumption increases more rapidly than with the Meyer type of 
valve. 

Summary.—The results of these tests are enlightening, in showing 
the actual amount of the losses occurring in the compression of air, 
particularly when the compressor is operating under the unfavorable 
conditions of varying air consumption necessarily obtaining in 


26 208 

e 

2 92 8 
a 

s 

AG 76s 
2 2 
=20 160.= 
2 Fs 
£18 144 > 
: ae { 
Fig eas OE 128 8 
RCT Na a 
Sig AN alee tl el he a 
8 Baca 2 
Sie %6 3 
: ia : 
5 0 80 
me eB 2 

8 — ene 
[as ae aSs = 5 

AGES Shae ee vas 
Hae 


32 
100 


Bt 
ES 
Goer 10 20 30 40 R Oe 60 10 80 90 


Fic. 125.—Test results of compressor plant No. 3. 


mining and other work in which machine drills play an important 
part. These losses are always recognized as existing by compressor 
builders and by intelligent users, and it is clearly desirable that 
properly conducted tests should be made more frequently. 

Again, compressor plants generally develop less power than their 
full rated capacity. It should be remembered that an air compressor 
is essentially a variable speed machine, its speed being regulated by 
some form of throttling governor, connected with the air-pressure 
regulator. The machine is therefore called on to run only as fast 


158 AIR COMPRESSION AND TRANSMISSION 


as the demand for air may require. It may be suggested that it 
would be well for compressor builders to give in their catalogues 
the actual horse-power rating at different speeds, with a table of 


ol RADSRGE Ea mE me 

SB laaaeimiee ieee aera ma a 
Bo A 

1600 


S 
oS 
Oo 


> 
oOo 
@o 
c=) 
o 


o 
35 © 4 600 


0 “10 pave 80 30 i000 


Fic. 126.—Test results of compressor plant No. 4. 


efficiencies at different loads and speeds, just as is done by some of 
the manufacturers of electrical machinery. Catalogues might also 
include some definite data respecting the cost per horse-power 
delivered by the air end of the compressor at different working 
speeds. 


CHAPTER XIV 


RECEIVERS. MEASUREMENT AND TRANSMISSION OF 
COMPRESSED AIR 


RECEIVERS 


The purpose of installing a receiver is four-fold: First, to equalize 
the pulsations in the air coming from the compressor; second, to 
collect the water and grease held in suspension by the compressed 
air as it leaves the compressor; third, to reduce the friction of air 
in the pipe system; and fourth, to cool the air as thoroughly as 
possible before entering the transmission system. 

It does not act primaiily as a reservoir of power, for in order to 
accomplish this its size would become impractical. However, in 


8 ‘ For 4'Satety 
=3 = ff Lhe | Valve ~, 
8 5 Lg) _ foe ». . 


TZ 


‘| 
WTAE 
4 i] 
| 
AWN AH 
HI Ht Hk Hl ie 
ay Hua H ' 
fa, pest SISISISESIS SISESISESISLSS 3 
ptt V9 its 
ghee at pS eR 


N 
a, 
eH 


LZZ 


ae, 4 
Lary) 


p~ 


Fic. 127.—Receiver aftercooler. 


compensating for the air pulsations it maintains constant pressure 
in the pipe line and in that way reduces friction. 

In order to facilitate the removal of water from the compressed 
air, it is frequently equipped with a coil of pipes (Fig. 127) filled with 
cooling water, in this way serving as an “after-cooler,” as it is 
called. When so equipped the difficulty with water in the trans- 
mission line and frost at the exhaust pipe of a compressed-air motor 
is reduced. 

When the pipe line is very long, receivers are placed at both ends 
of the pipe; this increases the effectiveness of the receiver and reduces 
materially the pipe friction. 

159 


160 AIR COMPRESSION AND TRANSMISSION 


As manufactured, these 1eceivers are usually supplied with a 
pressure-gage, safety-valve, blow-off cock and frequently a man-hole. 
They are made either horizontal or vertical and of cubical contents 
varying usually from 30 to 400 cu. ft. For exceptional cases as 
for compressed air-pressure water systems, they are frequently 
made much larger. 


THE MEASUREMENT OF AIR AND GASES 


‘The measurement of compressed air and gas in the commercial 
distribution and sale of these commodities and in testing com- 
F pressors has attracted a great deal of 

Seen te attention in recent times and excel- 

Re m ten pane Fe ul lent articles’ are to be found in the 
aa ci technical press. The material here 
given has been gathered from these 
sources and includes some interest- 
ing results of tests made in the 
Steam and Gas Engineering Labora- 
tory of the University of Wisconsin. 
“Standards of Measurement.— 
In making measurements it is usually 
necessary to ascertain the number 
of ‘standard cubic feet’ passing in a 
given time. The contents of a stan- 
dard cubic foot are determined by 
the assumed standards of tempera- 
ture and pressure used in defining 
the unit of measurement. Scientific 
data on gases are usually referred to 
the freezing temperature of water 
: and to the mean barometric pressure. 

Fic. 128.—Wet displacement Common commercial standards of 
meter. temperature and pressure in gas 
measurement are 60° F. and 30 in. 


jini 


a uideaidtidtin nn : y 
it i ce Md 


Wt 


of mercury, respectively. 

‘““A quite general classification of meters includes two main types: 
volumetric meters and velocity meters. 

‘Volumetric Meters.—Volumetric meters include what are known 
as “dry meters,’ operating on the general principle of a bellows, and 
‘wet meters.’ The latter are built in large sizes for use at gas works 


1The Measurement of Gases, Prof. Carl C. Thomas, Jour. Franklin Inst., 
Nov., 1911. Measurement of Nat. Gas, Thos. R. Weymouth, Jour. A. S. M. E.., 
Nov., 1912. Flow of Gas through Lines of Pipe, Forrest M. Towl, Lecture 
Columbia Univ., rort. 


MEASUREMENT OF COMPRESSED AIR 161 


in measuring the gas, as made, before being passed for storage to the 
holders (Fig. 128). These meters are known as station meters, 
their construction is, in general, that of a drum revolving within a 
cylinder or tank which is more than half filled with water. The 
revolving drum consists of a shaft carrying three or four partitions 
arranged in a spiral form. These partitions emerge in turn from the 
water as the shaft revolves, and each forms with the water a water- 
sealed compartment, which alternately receives and delivers gas. 
The drum receives its motion from the pressure of the gas itself and 
the number of revolutions of the shaft when properly calibrated give 
an index of the quantity of gas passing through the meter. 

In testing air compressors, volumetric methods of measuring the 
air compressed are sometimes used by installing three tanks. The 
compressor is arranged to discharge constantly into one of these 
at a constant pressure. This tank in turn discharges alternately 
into either of the other two. It fills one tank while the other is being 
discharged to the atmosphere and when the pressure approaches 
that of the compressor the discharge is turned into the empty tank. 
By noting the temperature and pressure and having the volume of 
the two tanks it is possible to calculate the volume of air which each 
has received from the compressor. 

“Velocity Meters.—Volumetric methods of measurement, how- 
ever, are not always feasible nor very satisfactory, and other methods 
of measurement depending on the velo- 
city of flow of the air or gas have been 
developed and made use of in commercial 
work. These methods may be separated 
into three types: the orifice or Pitot-tube 
type, which depends for its operation upon 
fundamental principles of hydraulics; the 
Venturi meter, which depends upon 
thermo-dynamic principles involved in the 
adiabatic expansion of the gas or air as it 
flows through the reduced cross-sectional 
area of the Venturi tube; and the heat 
meter, of which the Thomas electric meter, 
manufactured by the Cutler-Hammer Co., fy¢. 129.—Simple form of 
Milwaukee, Wis., is the best example, in Pitot tube. 
which the temperature of the gas or air is 
increased through a known range by a measurable amount of heat. 
From a knowledge of the specific heat of the gas and air, the weight 
of gas or air flowing through the meter is automatically determined 
and recorded. 


“Ditot Tube.—The Pitot tube (Fig. 129) affords a means of 
11 . 


162 AIR COMPRESSION AND TRANSMISSION 


measuling the velocity of air or gas through a pipe at any given 
point in the pipe section. In its simplest form it consists of two 
small tubes inserted in the pipe line—one having an opening pointed 
up-stream and communicating to one end of a U-tube the pressure 
due to velocity head in addition to the static pressure in the pipe; 
the other having an opening at right angles to the direction of flow 
and communicating to the opposite end of the U-tube the static 
pressure only. The difference between these two pressures is the 
pressure due to velocity alone, and from this, velocity of the gas or 
air in the pipe can be computed by means of the formula v?=2¢h 
where h is the static head necessary to give to the air or gas a veloc- 
ity of vft. persecond. From a knowledge of the cross-sectional area 


as Slots in Sides 
or Outer tke wane Z 
‘ ‘ LY 


Fic. 130.—Modern form of Pitot tube. 


of the pipe and the density of gas at observed pressure and tempera- 
ture, the quantity passing per unit of time can be computed. Fig. 
130 shows a modern type of Pitot tube. 

“The velocity of gas flowing through a pipe is not the same at 
all points in the section. It falls off gradually from the center out- 
ward and very rapidly near the inner skin of the pipe. In order to 
obtain accurate results with Pitot tubes, without exploring the pipe 
at several different depths, it is necessary to ‘standardize’ the tube 
and pipe together and find the depth at which the tube will indicate 
the mean velocity; that is, a Pitot tube will not necessarily give 
consistent readings if placed in a given position in pipes of different 
sizes, different conditions of surface, etc. The tube must be located 
with special reference to the size, shape, and condition of pipe with 
which it is used. Great care must be taken that the openings 


MEASUREMENT OF COMPRESSED AIR 163 


through which the pressures are communicated to the U-tube are 
properly placed with respect to the direction of flow, and they must 
be kept free from deposits. 

“The general formula for the Pitot tube and the orifice is derived 
from the law of falling bodies. Let 

T = absolute temperature of flowing gas, degrees Fahrenheit. 

P = absolute pressure of flowing gas, pounds per square inch. 

w = weight per cubic foot of flowing gas, at P and T. 

G = specific gravity of flowing gas, (air 1.0). 

v = actual velocity of flowing gas, feet per second. 

h; = height in feet of homogeneous column of gas at P and T 

producing 2. 3 

h = corresponding height of water column in inches. 

W. = weight per cubic foot of water, 62.37 lb. at 60° F. 

P, = absolute storage pressure base, pounds per square inch. 

[, = absolute storage temperature base, degrees Fahrenheit. 

Wa = weight per cubic feet air at 32° F. and 14.7lb.=0.08073 lb. 

d = actual inside diameter of pipe or orifice in inches. 

E = efficiency of Pitot tube or orifice. 

Q = flow in cubic feet per hour at P, and T,. 


Then 
V=A/2gh; =|2 8 a 
ie 
W=WaG aT 
vy ae pez 
OES ita Pal 
O= 218.44 ka? a 


“Prof. S. W. Robinson who was probably the first to use the 
Pitot tube in connection with the flow of natural gas has developed 
the following formula which has heen used by natural gas men for 
a number of years: | 


OQ =1,462,250 d? es) 0.29 os de 


This was derived from the formula for adiabatic flow 


n—1 
oo 2g X44 Po hf Pi\ on ie in which 
G1) w |G) geal 


164 AIR COMPRESSION AND TRANSMISSION 


v = velocity of flowing gas, feet per second. 

Py) = absolute pressure of the atmosphere, pounds per square 
inch. 

n = ratio of the specific heats. 

w = weight per cubic foot of gas at pressure P,. 

G = specific gravity of gas, air I. 

P, = absolute pressure shown by Pitot tube, pounds per square 
inch. 

d = internal diameter of well mouth, inches. 


= open-flow capacity of well, cubic feet per 24 hours. 
“Prof. Robinson has computed tables from the above formula 
which have been used for years. The computations are based on 
the following: 


nm =I1.408 
2g =04.3 
Ps =14.6 
Gir=0:0 


IE i i Se Rito Ne. 

To = absolute temperature of melting ice. 

T = absolute temperature of flowing gas. 

I’, = absolute temperature of storage. 

“Thos. R. Weymouth in his article in the Journ. A. S. M. E. 
points out that for natural gas the ratio of the specific heats is more 
nearly equal to 1.266 and by using 


Le OO nel. 
(le ree NE 
Vt ae ay) 


P, = storage pressure 14.65 
the formula becomes: 


if 0-21 a 
O=1,758,560 ay (44) Sh }* 
el AA 


“In order to obtain a mean value of / for the use in Pitot tube 
measurements Prof. G. J. Davis of the University of Wisconsin 
devised the following method which is illustrated in Figs. 131 and 
132 showing results of an actual test of a Pitot tube placed tandem 
with a Venturi meter and a Thomas electric meter. 

‘The horizontal represents distances from the center of the pipe 
at which readings of # were observed. On the vertical a suitable 
scale of values of \/h is laid off. Readings of ~/h are then plotted 
and joined by radial lines to the point representing the center of the 
pipe. The intersections of the slanting lines with the perpendiculars 


MEASUREMENT OF COMPRESSED AIR 165 


representing the positions at which the corresponding values of Vh 
were read are points through which a smooth curve can be drawn. 
The area under the curve may now be determined and from this 
the altitude of a triangle having the same area and base as the 
irregular figure will give the mean / to be used in computing the mean 
velocity. 

‘““The mean velocity V is determined from the formula V? =2¢h 
after reducing the # determined to equivalent feet of air. The 


pounds per hour will then equal 3,600 AVG where A is the area of 
the pipe in square feet. 


eee 

Gee 
A 
NAG 


\ EN 


LAREN GE aces 


ae 
rey 
: ! Val a aa BA 
I gt LA 
= ieee AA Sa VAC 
SSR Zasi ap aaa 
3 0.1 FAR =e 5 07 - Puta 
A oe. [eal nla 
Te ‘CCEA 
: FA “OOO sA 
} LG 04 J Zt 
ott wis =o L EL 
02 Eee 2a ae 02 Kt 
ale area ae eine it KAZ ie 
premanee Sent Center of bier eine Distance from Center oF Pipe, mohes, 
Fic. 131.—Graphical method of Fic. 132.—Graphical method of 
determining mean head. determining mean head. 


V is the velocity in feet per second. 

G is the weight of a cubic foot of air as it passed through the pipe. 

“‘In measuring air by means of a Pitot tube it is necessary to take 
into account the humidity of the atmosphere and make corrections 
as indicated in the discussion on Humidity given in Appendix C. 

‘‘In measuring large quantities of air in testing air compressors it 
is quite a common practice to have the air escape through a suitable 
orifice to the atmosphere. An apparatus of this kind is shown in 
Fig. 22 and one for large installations in Fig. 133. 


“The formula usually used for measuring air under these condi- 
tions is 


166 AIR COMPRESSION AND TRANSMISSION 


ie 
W =0.53A Tz where P; is greater than twice atmospheric 


pressure. 


7 TT, Tt UM, iets 
Ui ; 


/ “frien, "| 

; rT 7 ia ili > 

7 i 3 : eit ti! (Uy \ 
LT. TV. Ahi batons, Hil! 


(iil AR aig i ; \ 


NTR NNN aii AA |i NN 


a : a in i ' al 
gu a + Ge wy gulls ei 


Fic. 133.—Apparatus for measuring large quantities of air. 


\ 


When P, is less than twice atmospheric pressure the formula usually 


used is 
W =1.060A A eeieee 
163 


“This last formula, however, is 
A not entirely reliable (see ‘Air 
Flowing into Atmosphere through 
Circular Orifices’ by R. J. Durley, 
Trans, ASS. MAES Y ol27 
In the above formule 
W = weight of air escaping in 
pounds per second. 
P. = pressure of atmosphere in 
pounds per square inch. 
P, = pressure before the nozzle 
in pounds per square inch 
absolute. 


(NNhhawm vrssr aa seees. 


4 T, = absolute temperature of 
SI] Ss . ° 
NIN air entering the nozzle. 
A = area of nozzle in square 


foc 
“Tn using a nozzle or orifice it 
is also necessary to consider the 
humidity of the atmosphere in 
measuring air. 
“St, Johns Meter.—A number of meters have been made making 
use of an orifice for measuring the flow of air. Such meters are usu- 


PRaveeaaaarvees 


Fic. 134.—St. John meter. 


MEASUREMENT OF COMPRESSED AIR 167 


ally calibrated by means of a gasometer. The St. Johns meter, 
Fig. 134, is in effect a variable orifice meter. The position of the 
plug S determines the size of the orifice through which the air passes 
and a graphical record is kept of the position of this plug on a drum 
moved by clock work and by planimetering this chart the average 
position can be determined and the consumption be calculated. 
‘Venturi Meter.—The Venturi meter, Fig. 135, consists of a throat 
or gradually contracted portion of the passage, which causes a de- 


ik 


{AU IOOU MOOI 


Fic. 135.—Venturi tube. 


crease in pressure and increase in velocity of the gas flowing through 

it. 

area of the up-stream section in square feet. 

Az = area of throat in square feet. 

P, = pressure at up-stream side, pounds per square inch. 

P., = pressure at throat, pounds per square inch. 

G, = weight of gas at up-stream section, pounds per second. 

n = ratio of specific heats, constant pressure to constant 
volume. 

V2 = velocity of gas at throat, feet per second. 


te 
O 
o 
aN 
I 


“By equating the loss in potential energy to the increase in kinetic 
energy it is found that 


N 
bo 
| 
Pgs 
a 
Saas 
ty] 
Pt aa 
to 
oR 
S 
eer 2 
NH 
= 
| 
ST ee 
| 
a 
| 3 
| 
Lon] 
dole 


pe 4a \7(£2)\2 
ae AyIEN Ps 
1 


ce . ° ie ee ° 
The quantity flowing O=A2V.G1 (3) ” in cubic feet per second. 
1 


“Tt is frequently necessary to take small readings of pressures 
with both the Pitot tube and Venturi meter, and in order to do this 


168 AIR COMPRESSION AND TRANSMISSION 


accurately the water columns should be read with a micrometer 
gage or differential (inclined) water column. 
‘“‘A similar formula for the flow expressed in cubic feet per hour 


would be 
Po ned 
Ts n Jey ‘poe ; Gal 
= O40 Ag) ee eee 2\ in 
Cee PNG JT; (>,) 


“Terms in this formula not appearing in the other are 

Q1 = flow in cubic feet per hour. 

T, = absolute temperature of gas at entrance. 

I, = absolute temperature of storage base pounds per square inch. 

P, = absolute pressure of storage base pounds, per square inch. 

G = specific gravity of gas, air=1. 

“Thomas Meter.—The Thomas electric meter is based upon the 
principle of heating the air or gas through a known range of temper- 


y L <p SS SS 


PASS aN [ASS 


sy 


SAA COLL ALLL TOTO ly Zz y 


Gee we beer rr} 


Fic. 136.—Sectional view of Thomas electric meter. 


ature. This measured energy is proportional to the weight of the 
gas flowing. Electric energy is used as a source of heat because 
it can be so conveniently and accurately applied and measured. 
Electrical resistance thermometers are used to regulate the tem- 
perature range through which the gas is heated, because with ther- 
mometers of such type very small differences of temperature can be 
accurately and easily determined. 

‘““A sectional view of a ro-in. meter and casing is shown in 
Fig. 136. Fig. 137 shows the meter diagramatically. The electric 
heater is placed within the casing between two electric resistance 
thermometers, 7; and 7». The heater consists of spiral turns of 
bare nichrome resistance wire wound around a conical frame and 
supported by insulators, so that heat is dissipated evenly over the 
section of the pipe. A rheostat is placed in the heater circuit for 
regulating the direct current supplied. This energy is measured 
by an ammeter and voltmeter. 

“The thermometers consist of resistance wire wound on wooden- 


MEASUREMENT OF COMPRESSED AIR 169 


spindles and evenly distributed over the section of the casing. The 
wire is of such material that its resistance increases with the tempera- 
ture according to known laws. The two thermometers form two 
arms of a Wheatstone bridge, the other two arms being fixed coils 
of wire that have a zero temperature coefficient. A galvanometer 
is connected across the Wheatstone bridge, and a small rheostat is 
placed in series with one thermometer for balancing the bridge when 
no heat is passing through the heater. A small resistance R; is 
arranged so that it can be placed in or out of series with the entrance 
thermometer. ‘This resistance is equal in value to the increase in 


Balancir 
rheosn 


y Pipe 
ie w 
Hew r Ley 
aS Se 
peel le RS i\/| §38 
Direction of 2 S ESS 
Flow SSE Xss 
S | Heater x 
ie aN | 
Water 
( g Rheostay 


Ammeter 


S 
> 
3 
\. 
iS 


| 


Fic. 137.—Diagrammatic sketch of Thomas electric meter. 


resistance of the exit thermometer for a rise in temperature of prac- 
tically 2° F. The meter in the laboratory of the University 
of Wisconsin is for 2.0152° F. 

‘The operation of the meter is as follows: With gas flowing through 
the meter but with no energy in the heater, and with R; out of cir- 
cuit, the two thermometers are brought to the same balance by 
means of the balancing rheostat and the galvanometer. Then the 
resistance R: is put in circuit and sufficient electrical energy is 
supplied to the heater to bring the galvanometer to balance again, 
by bringing the exit gas to a temperature 2.0152°, with the meter 
mentioned, higher than that of the entering gas. The measuring 
instruments in the heater circuit then indicate the energy required 


170 AIR COMPRESSION AND TRANSMISSION 


to raise the temperature of the air or gas through a known range. 
The quantity of gas flowing can be found by the equation 


Ww a 3412k 
ts 


where W is the number of pounds of gas or air per hour, # the amount 
of energy supplied in watts per hour, ¢ the rise of temperature in 
degrees Fahrenheit, and s the specific heat at constant pressure of 
the gas or air. 

“With the laboratory meter the air flowing through the meter 
per minute is given by the formula 0.028218 E;. 

“In applying this meter to gases it is necessary to ascertain the 
composition of the gases in order to obtain the mean specific heat 
for use with the meter. 

“The meter in commercial form is equipped with automatic 
devices to regulate the flow of current through the heater so as to 
maintain a constant difference of temperature between the resistance ° 
thermometers of 2°. The electrical instruments for measuring the 
consumption of current in the heater are then calibrated to read either 
weight or quantity of gas flowing and this reading is recorded 
graphically. 

“Meter Comparisons.—At the University of Wisconsin tests 
were run by placing a Venturi meter, Pitot tube and Thomas meter 


SAG To ‘ 
Sweet 
Pitot TORE fom Atte 


0 
Driven Fan 


Fic. 138.—Sketch of meters placed tandem for testing. 


in tandem, as shown by Fig. 138. The results of these tests are shown 
as Fig. 139. A remarkable similar set of readings were secured. 
“In April, 1911, a Thomas meter was tested on a natural-gas line 
by comparison with Pitot-tube measurements giving practically iden- 
tical results. This meter had a maximum capacity of 750,000 cu. ft. 
of free gas per hour and an accurate minimum capacity of 12,500 cu. 


MEASUREMENT OF COMPRESSED AIR 171 


ft. It gave a continuous graphical record and integrated values of 
the gas directly in standard cubic feet at 15.025 lb. absolute pressure 
and 60° F., although the pressure of the gas varies from 50 to 200 |b. 
gage and the temperature varies with weather conditions. The 
specific heat was calculated from an average analysis of the gas for 
the standard conditions given above. This particular meter was 
placed in a ro-in. line and located about a mile and a half from a very 


ena, co 
+ Scere ts 


Gas Engineering ee 
University of Wisconsin, June 2-/9/1 


Electric Meter « 
Venturi Meter x ——— 
Pitot Tube o—---- 


0 
0 100 200 300 400 500 600 700 800 900 1000 1100 
R.p.M. of Fan 


Fic. 139.—Result of test. 


complete Pitot-tube meter station. A 22-hour comparative test 
showed a difference of 0.2 per cent. between the two meters and a 
similar comparative test from April 17 to June 3, 1911, showed the 
same difference.”’ 


_ PIPE LINES! 


“The transporting of gas or air requires a line which shall be “air 
tight.” It is much more difficult to make a line to hold gas or air 
under pressure than it is to hold a liquid. Trouble has been expe- 
rienced in almost all lines built for high pressure on account of the 
leaking of the gas at the couplings. The first high-pressure lines 
were laid with bell and spigot joints, caulked with lead. The lines 


1 Forrest M. Towl, Lecture Columbia University, 1o1t1. 


siz AIR COMPRESSION AND TRANSMISSION 


might be tight when they were first laid but the movement in expand- 
ing and contracting soon caused them to leak in large amounts. 

‘The next lines used were of wrought iron or steel pipe with screw 
joints. While these held much better than the bell and spigot type, 
there was still enough leakage to make it desirable to have a more 
perfect joint. The leakage on some of the earlier screw-joint gas 
lines was such that by putting a rubber bag over the coupling, gas 
could often be collected at the rate of from 20 to 50 cu. ft. per hour, 
or enough to run a good-sized torch. This was true of lines up to 
8 or ro in. in diameter. When the lines became larger, the leakage 
increased so much that it was practically impossible to use large 
size lines and get a large percentage of the product at the market. 

‘‘As the demand for natural gas increased, it became necessary 
to use larger lines, and a rubber packed stuffing-box was developed. 
The first successful joint of this kind in the market was the Dresser 
coupler, and it is due largely to this and other couplings that the 
natural-gas industry has become so great. | 

‘Dresser Coupler.—The Dresser coupler consists of a sleeve into 
which the ends of the pipe are placed. There is a projection at the 
center of the sleeve so that the ends of the pipe 
will be each inserted into the sleeve the same 
distance. This sleeve acts as a follower to 
compress rubber in an annular space into the 
end rings which are drawn together by bolts. 
The rubber is surrounded on one side by the 
pipe, on another by the body of the coupling, and 
Fic. 140.—Dresser On the remaining side by the end rings so that 

pipe coupler. there is very little of the surface of the rubber 
exposed either to the gas on the inside or the 

air on the outside of the line. It is found that these joints will last 
for years. (Fig. 140 shows a cross-section of the Dresser coupler.) 

‘‘“Hammon Coupler.—The Hammon coupler is a modification of 
the Dresser, one of the principal features of which is that the pro- 
jection at the center of the sleeve is made by lugs welded onto the 
sleeve. When it becomes necessary to take apart one of these 
couplers, the lugs can be broken off and the coupler slipped back so 
as to allow the pipe to be easily removed. (Fig. 141 shows the Ham- 
mon coupler.) 

‘Lines of pipe can be built in almost any kind of country, but it 
is necessary in some places to arrange to keep the lines from acting 
as a Bourbon tube and expanding in one direction until the ends 
of the pipe may be pulled out of the coupling. To avoid this trouble 
it is customary in such places as river crossings to use screw pipe, 
and to place over the collar a clamp which is constructed to make a 
rubber joint between the ends of the collar and the pipe. 


\ 
) } tt 

iain 

PS SS SSS SB 

RS ca 

ereveereeenrrnrnrernrrrrrr rr 


MEASUREMENT OF COMPRESSED AIR 173 


“For power-transmission lines or for temporary gas lines, where 
the distances are short or the service temporary, it is not considered 
necessary to bury the pipe, it will be found that the screw-joint pipe 
is satisfactory, but for other 
natural-gas or air service, the 
rubber coupling has many things 
to recommend it, and when the 
capacity requires large pipe, it is 
almost absolutely necessary to use 
this type of coupling. These 
coupling have been used for manu- 
factured gas, but it is found that 
the condensation from the gas 
collects in the coupling and soon 
causes a leak in the rubber joint. 
Work is now in progress to per- 
fect a material which will not be 
acted upon by the condensation 
in the gas and which will make a . 
gas-tight joint. Fic. 141.—Hammon pipe coupler. 

In using screwed joints for air 
it is necessary that the lead, litharge, or other material used at the 
joint should be applied on the ends of the pipe and not in the coup- 
lings, so that the surplus is brought outside instead of within the 
pipe where it may cause a more or less serious obstruction.” 

Pipe-line Formulz.—A very simple formula is often used for 
calculating pipe lines for compressed air. 


Deo a) or pee pi~ bs 
Wy V/] Wi 


in which 

D =the volume of compressed air in cubic feet per minute dis- 
charged at the final pressure. 

¢ =a coefficient varying with the diameter of tke pipe, as 
determined by experiment, 

d =diameter of pipe in inches, 

1 =length of pipe in feet, 

pi =initial gage pressure in pounds per square inch, 

p2 =final gage pressure in pounds per square inch, 


1 The actual diameters of wrought-iron pipe are not the same as the nomina- 
diameters for all sizes. This difference is small, however, except in the 1 1/4 in. 
and 1 1/2 in. sizes, the actual diameters of which are 1.38 in. and 1.61 in. 
respectively. 


174 AIR COMPRESSION AND TRANSMISSION 


w, =the density of the air, or its weight in pounds per cubic foot 
at the initial pressure #1. 
The second form of the formula, as given above, will be found 
convenient for most calculations, as the facto1s can be considered in 


groups. 


In Tables XIII and XIV are given the values of c, d°, and 
c\/d5, The values of c show some apparent discrepancy for sizes 
of pipe larger than 9 in. but there would be no very material dif- 
ferences in the results. 


» TABLE XIII 

Diameter of pipe, Values of Fifth powers of Values of 

inches C d c\/ds 
I 45-3 i 45-3 

2 5200 a2 207 

3 56.5 243 876 

4 58.0 1,024 1,856 

5 59.0 3,125 3,298 

6 59.8 7,776 5,273 

7 60.3 16,807 7,317 

8 60.7 32,768 10,988 

9 61.0 59,049 14,812 

IO 61/32 100,000 19,480 

i 61.8 161,051 24,800 

12 62.0 248,832 30,926 


TABLE XIV.—VALUES OF Wi FOR ET ACE EGS pene UP 100 LBS.PER SQUARE 


Gage pressure, : | — Gage pressure, = 

ap a Vw | pounds - Vw 

° 0.0761 0.276 “is 0.3607 0.600 

5 0.1020 0.319 60 0. 3866 0.622 

ite) 0.1278 0.358 65 eg As 0.642 

cf On 527 0.302 70 0.4383 0.662 

20 0.1796 0.424 7 o.4042 0.681 

25 0.2055 0.453 80 0.4901 0.700 

30 O. 2313 0.481 85 0.5160 0.718 

25 O.2572 0,507 go 0.5418 Of736 

40 0. 2831 0.532 95 0.5677 0.753 

A5 ‘0.3090 0.556 Too 0.5936 0.770 
50 0.3348 O. 578.2 eva 2 yan a he he al cea OR te ee 


MEASUREMENT OF COMPRESSED AIR 175 


Mr. Frank Richards gives the following formula for determin- 
ing the loss of pressure in pipes: 


2h 


10,000D°a 


from which 


10,000D*a 


L 


V= 


In these equations 
D=diameter of pipe in inches. 
L =\length of pipe in feet. 
V =volume of compressed air delivered in cubic feet per min. 
H =head of difference of pressure required to overcome friction 
and maintain the flow. | 
a =constant depending on the diameter of the pipe. 


TABLE XV.—VALUES OF a, D' AND Dia FOR WROUGHT-IRON PIPE. 


Gane ; Ds Dia 
pipe diameter 
I in. 0135 I 023% 
1% in fans RES Ta52s 
Tz in 0.662 7.59 5,03 
ain, 0.565 ex 18.08 
anit 0.65 97.65 63.47 
3 in 0.73 243. ETA 
33 in O87 525. 413.2 
4 in 0.84 1024. 860.2 
5 in 0.934 2125. 2010075 
6 in I .000 vie by doe Os 
8 in Te ks 32768. 36864. 
ro in pane 100000. 120000. 
2-10 T20 248832. 213525. 
16 in N34 1048575. I4O5001. 
20 in 1.4 3200000. 4480000. 
24 in I.45 7962624. 11545805. 


For example, suppose it is desired to determine the loss in 


pressure in transmitting 300 cu. ft. of compressed air per min. 
through a 6-in. pipe one mile in length. 


L =5280, D*a for a 6-in. pipe=7776 
2 
Nehey vind eb EBON Ne) 


10000 X 7776 
6.11 lb. 


That is, the pressure drop will be 


176 AIR COMPRESSION AND TRANSMISSION 


As another example, suppose it is desired to ascertain the 
proper size of wrought-iron pipe for transmitting compressed air 
from a compressor of 1500 cu. ft. free air capacity per min. at 
80 lb. gauge a distance of 2000 ft., with an allowable loss of 
pressure of 5 lb. 

The pressure at delivery will be 75 lb. gauge or. practically 6 
atmospheres. ‘The volume of compressed air delivered per minute 
will be: 


1500+6=250 cu. ft. per min. =V 


As H=5 the formula 
Paes 

tooo0oH 
250” X 2000 

10000 X 5 ere 

From Table XV it is seen that D*a for a 5-in. pipe is 2918.75 
and for a 6-in. pipe 7776. This would indicate the advisability of 
selecting 5-in. pipe for the conditions of this problem. 

The friction in pipe elbows may be expressed in terms of 
equivalent lengths of straight pipe. Elbows having the largest 
radius will naturally give the least friction and the accompanying 
table as given by the Norwalk Compressor Co. gives the friction 
effect of elbows in terms of the radius. 


may be used with proper substitutions, from which 


5 


TABLE XVI.—FRICTION EFFECT OF ELBOWS IN TERMS OF PIPE LENGTHS 


Radius of elbow in 
pipe diameters 


On 
Ww 
bo 
eS 
dle 
eS 
| 
vol 
ated 
[<i 


Equivalent lengths of straight 


pipe in pipe diameters 7.85 |8.24 9.03 |10.36/12.72/17.51|35.09 leas 


REHEATING 


From a consideration of changes that take place during the 
compression and expansion of the air, it is apparent that heating the 
air just before expansion will raise its temperature and impart to it 
an increase of energy which, if used immediately, will increase the 
efficiency of the compressed air. In addition, this reheating will 
increase the temperature at the end of expansion and prevent the 
particles of moisture in the air from freezing. 

It is not uncommon in the ordinary use of compressed air to find 


MEASUREMENT OF COMPRESSED AIR aT 


exhaust temperature varying from 5° to60° F. The lower tempera- 
ture is very apt to cause trouble particularly in out-door work 
during the winter months. It is quite probable that reheating was 


5 FL =" * 
1C) (|| 4Inler 
Sen 


p 
Soe) — 
ieneatareeeseneeeny = 5 
4 
Ns s 
— iy 
f S 
a) 
4 kay 
i 
(i 


RRS ones 
verti T3054 
RRR Rr fae 


POOR 

SER 
eres 

SIV 


PSR 
H IRS 
pe 


Reheater Coil 


Fic. 143.—Leyner air reheater. 


first introduced to prevent the formation of frost in the exhaust pipe 
and its advantages are so apparent that no economical use of com- 
pressed air in large installations is complete without some system of 
reheating. 
Stoves.—In quarry work it is very common to make use of stoves, 
12 


178 AIR COMPRESSION AND TRANSMISSION 


such as are shown in Figs. 142 and 143, for heating the air just before 
its use in the drills. In locomotive work for mines and surface use, 
hot water is frequently used to heat the air. <A recent locomotive 
built for the government for handling cars of explosives has a com- 
pound cylinder, expansion taking place in one to such a point as to 
bring the temperature below that of the atmosphere. ‘The air then 
passes to the low-pressure cylinder through pipes in contact with the 
atmosphere and in this way its temperature and energy are increased. 

Reheaters are usually capable of raising the temperature of the air 
to from 300 to 500° F., although common practice shows temperatures 
from 250 to 350°. In figuring on reheaters it is usual to assume that 
t lb. of coal will give from 8,000 to 10,000 B.t.u. to the air. As the 
specific heat is 0.0686 B.t.u., it is evident that 1 lb. of coal will heat 
33 lb. of air or about 408 cu. ft. to a final temperature of 350°. As 
reheating increases the volume as well as the temperature, the econ- 
omy in its use expansively is quite evident. 


CHAPTER XV 


THE SELECTION AND CARE OF AIR COMPRESSORS 


It is difficult, if not impossible, to dictate the type of compressor 
that must be selected to suit certain conditions, as the importance 
of the various factors influencing the selection vary greatly in differ- 
ent localities; however, a brief statement of some of the factors to be 
considered may be of assistance. 

Available Power.—The character of driving power available will 
usually determine the type of prime mover to be selected. The 
principal manufacturers of compressors have on the market the 
belt-driven, steam-driven, and electrical-driven machines, as well 
as those driven by gas engines or water turbines. 

For small and intermittent demands it is difficult to make use of 
an economical steam engine for operating the compressor, and for 
‘this reason when belt power is available and suitable for the condi- 

tions of operation, the belt-driven compressor is to be preferred. 

When the quantity of air to be compressed is large, it will usually 
be found advisable to install a steam or motor-driven compressor as 
this allows greater flexibility and economy of operation. 

If this type of motive power is selected, a choice of different de- 
signs must be made. The straight-line type, equipped with proper 
energy-compensating devices to secure economy of steam consump- 
tion, has the advantage of simplicity and high mechanical efficiency. 
On the other hand, the duplex type appeals to many, for, if 
single-stage, it may be installed in sections and in this way future 
extensions can be made with minimum expense. 

The pressures that are desired will determine whether the com- 
pressor is to be of the one-, two- or three-stage system; the first being 
usually selected for pressures below 80 lb. per square inch, the second 
for pressures from 80 to 500 lb., and the third for pressures from 
500 to 1,000 lb. per square inch. 

Valve Gear.—The price of fuel will influence largely the type of 
‘valve gear to be selected for a steam engine. If fuel is very cheap 
and its consumption comparatively unimportant, the simpler forms 
of valve gears are to be preferred. If, however, the economy of 

179 


180 AIR COMPRESSION AND TRANSMISSION 


fuel is of importance, the more complicated and expensive types, 
such as the Corliss, should be selected. 

The plain slide valve, the independent cut-off valve, such as the 
Meyer, and the Corliss give steam consumptions which decrease 
in the order named. The last two types, particularly the Meyer, 
are quite common for steam-driven compressors. 

In the mountainous sections of the country and in those parts 
where the supply of water-power is abundant, water-wheels or 
turbines are largely used as prime movers. 

The distance.from the source of power to the place where the 
compressed air is to be used will determine whether the water- 
wheel is to be coupled directly to a compressor or to an electrical 
generator which will generate current to be transmitted to an 
electrically driven compressor. 

Both types are in use in the mining regions, some of the largest 
compressed-air installations being equipped with compressors 
driven directly by water wheels. 

Within the last few years the method of compressing air by means 
of a waterfall without the use of any mechanical parts has increased 
to such proportions as to demand the attention of engineers con- 
nected with the installing of compressed-air equipments. 

Air compressors driven by gas or gasoline engines are frequently 
used in quarries and other places where it may be desirable to move 
the compressing plant from point to point. Where gas can be 
obtained cheaply it makes a most desirable machine, because of 
the high efficiency of the gas engine. Within the last few years 
gas engines operated by the gases from blast furnaces have been 
developed to such an extent that many of the large blowing engines 
for Bessemer converters are operated in this way, giving compressed 
air to the converters and other places where used at a minimum 
expense. . 

In selecting any type of air compressor, particular attention 
should be paid to the construction and design of the valves. 
Mechavically operated inlet and automatically operated discharge 
valves seem at present to represent the favorite practice, although 
automatically operated inlet valves are preferred by many 
because of the little attention they require. 

All valves should be simple in construction, of large port opening, 
durable and reliable in action, and easily removed for purposes 
of examination and renewal. 

The largest possible amount of surface, including cylinder heads, 


SELECTION AND CARE OF AIR COMPRESSORS 181 


should be water-jacketed on all piston compressors except those 
discharging at very low pressures. 

The nature of the work performed by the compressor will deter- 
mine the advisability of installing an unloading or governing device, 
economy of operation usually demanding the installation of some 
such apparatus. 

Size and Type of Compressor.—In order to give an idea of the 
data required when determining the size and type of a compressor 
to be selected, the following is given, as published by the Sullivan 
Machinery Company in their catalogue: 

1. Purpose for which the compressed air is to be used (coal- 
mining machines, rock drills, air lift, etc.). 

2. Volume of free air required in cubic feet per minute. 

3. Working air pressure. 

4. Altitude at which compressor will work, if over 1,000 ft. 
above the sea-level. 

5. Number, size, and class of machines to be operated by the 
compressed air. 

6. If the air is to be used for pumps, give make, size, and speed 
of pump and height to which water must be delivered. 

7. If for raising water by the “‘air lift,”’ state desired flow per 
minute in gallons, diameter and depth of well, and height to which 
water must be delivered, measuring the average height of water 
in the well. 

8. Will the demand for air be constant or intermittent during 
the daily time of operation? 

9. Will the compressor be operated by steam or power? 

10. If steam driven, state working steam pressure, kind and 
average cost of fuel available, type of engine preferred and whether 
it is to be run condensing or non-condensing. 

tr. If power driven, state motive power (as water-power, 
electricity, rope driven, or gasoline engine) and whether direct 
connection, belt, or gearing is preferred. 

12. If water-power is to be used, give horse-power available, 
or head or fall of water in feet, also amount of water supply in cubic 
feet per minute. 

13. If belt drive is employed, give horse-power at belt. 

14. State facilities for transporting compressor to destination. 
If machine must be sectionalized, state means of transportation, 
heaviest weight allowable for a single package and number of pack- 
ages permissible of maximum weight. 


182 AIR COMPRESSION AND TRANSMISSION 


COMPRESSED AIR EXPLOSIONS 


‘“‘Compressed air claims to be and is a safe power. Occasionally 
we hear of a case of firing, which to some may appear to be a serious 
objection to the use of air; but if the causes are known and under- 
stood and due care is observed, firing becomes a matter of care- 
lessness. Compressed air is not inflammable, but, during com- 
pression by piston compressors it is necessary to use oil for 
lubrication, and this oil or the gases from it form a combustible 
mixture with the air. 

In most cases firing may be traced to the use of poor oil, but in 
others too much oil sometimes causes ignition. 

Lubricating Compressors.—It is a common mistake of engineers 
to feed oil too rapidly to the air cylinders. A drop now and then 
is all that is required to keep the parts lubricated. The air cylinder 
does not require as much lubrication as the steam cylinder, for there 
is no tendency to cut and wash away the oil as there is in a steam 
cylinder. 

When too much oil is used, there is a gradual accumulation of 
carbon which interferes with the free movement of the valves and 
which chokes the passages, so that a high temperature may for a 
moment be formed and ignition follow. 

It is well to get the best oil and use but little of it. 

There are cases where firing has arisen from the introduction of 
kerosene or naphtha into the air cylinder for the purpose of cleaning 
the valves and cutting away the carbon deposits. This is a very 
effective way of cleaning valves and pipes, but it is a source of danger 
and should be absolutely forbidden. 

The inflammability of benzine, naphtha and kerosene is so 
acute that it is a dangerous experiment to introduce anything of this 
kind into an air cylinder. 

Cleaning Valves.—Soft soap and water is the best cleanser for the 
air cylinder and it is recommended even in cases where the best oil 
is used and it is a good plan to fill the oil cup with soft soap and 
water and feed it into the cylinder, as the oil is fed, at least once or 
twice a week, or even oftener if necessary, in order to prevent the 
carbon deposit from gumming up the valves. 

A thick or cheap grade of cylinder oil should never be used in an 
air compressor. Thin oil which has a high flash-point and which 
is as free from carbon as conditions of lubrication will admit is the 
best oil. 

There may be considerable danger in a valve which is so gummed 


PEEL CIIONGAND CAKE OPSATR COMPRESSORS, 188 


as to be unable to close at the right time. When a piston has 
compressed air and forced it through the discharge valve and then 
starts on its return stroke, there is immediately a tendency for 
the air just compressed to return to the cylinder, and if the valve 
does not operate properly there will be some hot compressed. air 
in the cylinder when the piston starts again on its compression stroke 
with air at an initial temperature, 200° or 300° above the normal. 
The final temperature at the end of compression will in consequence 
be quite high and may even be above the ignition point of the lubri- 
cating oil that is used. 

To guard against this, care should be taken in the selection of the 
type of compressor used and the valves and passages should be 
thoroughly cleansed once a week by the engineer, who should also 
investigate the valve seats to insure the valve fitting properly when 
in place. 

Inlet Connection.—The inlet should be closed in a cold-air box, 
or some direct connection should be made to the outside air in order 
to avoid taking in hot air to the compressor. 

In compressors used in the coal-mining districts, care should be 
taken to see that coal-dust is not drawn into the cylinder. 

A thermometer should be placed in the discharge pipe close to the 
compressor so that the operating engineer may note any change of 
temperature and stop or slow down the compressor to avoid an 
accident. 


APPENDIX A 


Common Logarithms.—The following table of common logarithms 
will be found of assistance in solving work and power problems 
dealing with compressed air. 

The table shows the “mantissa”? only, the ‘“‘characteristic”’ 
depending on the location of the decimal point and being one less 
than the number of figures to the left of the decimal point of 
the given number. For example: 

The logarithm of 529 is 2.7235 

The logarithm of 52.9 is 1.7235 

The logarithm of 5.29 is 0.7235 

The logarithm of 0.529 is 9.7235—I10 

The logarithm of 0.0529 is 8.7235 —Io. 

The table of proportional parts is for the purpose of interpolation. 
For example, if the log of 80.54 is desired it is found as follows: The 
log of 80.5 from the table is 1.9058. The table of proportional parts 
shows in the same line of figures the additional figure to be added 
under the column marked 4 to be 2, making the required log of 80.54 
1.9058 plus 0.0002 or 1.9060. 

Logs of powers of numbers are found by multiplying the log of 
the number by the given power or exponent. For example, suppose 
it is required to find the value of 28.31". 

The log of 28.3 is 1.4518, and this multiplied by 1.4 is 2.0325, that 
1s, the logrol 2823" 2"se290375. 

The antilog of this or the numerical value of 28.31:4.is found by 
looking in the table for the number whose logarithm has a “‘mantissa”’ 
of 0325 and then pointing off three places from the left or one more 
than the characteristic 2. The antilog of 2.0294 from the table 
is 107, and as the given log is 0.0031 higher than 2.0294 the table 
of proportional parts would indicate that the antilog of 2.0325 is 
about 0.75 higher than 107. That is 28.31:4 is equal to 107.75. 

The principal difficulty in handling logarithms of small numbers 
with fractional exponents is met in dealing with the characteristics. 
This may be treated as follows: 

Suppose the value of 0.483°°* is desired. 

The log of 0.483 is 9.6839—10 

184 


APPENDIX A 185 


LOGARITHMS. 


Proportional Parts. 


8 
12 3/4 5 


0043|0086|0128 
0453)0492/0531 
0828/0864/0899 
1173)1206)1239 
1492)1523/1553 


0212/0253 
0607|0645 
0969}1004 
1303}1335 
1614/1644 


0334 
0719 
1072 
1399 
1703 


-— 
= 


(SU SCOR at 
& &~100 
ht pe 
BSS 
— ee 
bo ow 


— 
— 


1790/1818) 1847 
2068) 2095/2122 
2330| 2355/2380 
2577/2601) 2625 
2810/2833) 2856 
3032|3054/3075 
3243/ 3263/3284 
3444/3464/3483 
3636|3655|3674 
3820/ 3838/3856 


1903/1931 
2175/2201 
2430/2455 
2672) 2695 
2900) 2923 
3096/3118) 3139 
3304/3324/3345 
3902/3522/3541 
3692/371113729 
3874} 3892/3909 


1987 
2253 
2504 
2742 
2967 
3181 
3385 
3579 
3766 
3945 


— 
—_ 


bo bo bo 09 
He OT OT OI Od 
~I ~I ~1 00 00 
— 
o 


OU D> OH HD 


4048) 4065/4082 
4216)4232|4249 
4378/4393|4409 
4533/4548) 4564 
| 4683)4698/4713 
4857 
4997 
5119/5132 
5250/5263 
5366!5378/5391 


4116 
4281 
4440 
4594 
4742 
4886 
5024 
5159 
5289 
5416 


3997/4014 
41664183 
4330/4346 
4487/4502 
4639) 4654 


m bo bo bO bO bo bo bo bo bo 
wwwwo PRP PA 


He Or Or Or Or 


4786/4800|4814 
4928) 4942/4955 
5065/5079] 5092 
5198}5211/5224 
5328/5340) 5353 


ee 
Oona “I CO 00 CO © 
CO 00 00 00 CO 


bo bo BO bO bo Co WO OO 


5490|5502/5514 
5611}5623/5635 
5729 5740|5752 
5843/5855|5866 
5955/5966)5977 
6064) 6075) 6085 
6170/6180/6191 
6274/ 6284/6294 
6375)/6385|6395 
6474/ 6484/6493 


5539 
5658 
5775 
5888 
5999 
6107 
6212 
6314 
6415 
6513 


5453/5465|5478 
5575/5587|5599 
5694|5705|5717 
5809| 5821/5832 
5922/5933|/5944 


Www, > eee PP 


See et 
CLD DAN 
NIT I 


6031)|6042|6053 
6138/6149/6160 
6243/6253/6263 
6345) 6355/6365 
6444/6454) 6464 


a 
Co 0) 09 09 OO 


6609 
6702 
6794 
6884 
6972 
7059 
7143 


6580)6590 
6675/6684 
6767|6776 
6857|6866 
6946/6955 
7033|7042 
7118|7126 


6542|6551/6561 
6637| 6646/6656 
6730|6739|6749 
6821/6830|/6839 
6911/6920/6928 
6998) 7007/7016 
7084/7093'7101 


ee 
Drrwryrmp nwwnwpnywrn 
09 09 Oo Oo 09 

RNS No, Mon Moy Mo ea eg en 
AMARA ARMRAO 
AOMINTN NIWA 
00 00 00 00 © 


| 


| 


7168/7177 
7251/7259 
7332/7340 


7185 
7267 
7348 


7202 
7284 


7364 


7210 
7292 
7372 


7226 
7308 


7388 


| 
bo bo bo bo bo 


bo bo bo © CO 
Re ee 
Or Or Or Or Or 


C2 2 2 > 


AOon4nna “I ~I ~I “100 00 CO 00 CO © 


~I ~I ~1 00 00 


186 


AIR COMPRESSION AND TRANSMISSION 


1 2 


7419 
7497 
7574 
7649 
7723 


7796 
7868 
7938 
8000/8007 
8069) 8075 
8136/8142 
8202/8209 
8267|8274 
8331)8338 
8395)/8401 


7412 
7490 
7566 
7642 
7716 


1789 
7860 
7931 


8463 
8525 
8585 
8639/8645 
8698/8704 
8756/8762 
8814/8820 
8871'8876 
8927/8932 
8982/8987 


8457 
8519 
8579 


9042 
9096 


9036 
9090 
9143/9149 
9196/9201 
9248/9253 
9299/9304 
9350/9355 
9400/9405 
9450/9455 
9499/9504 


9552 
9600 
9647 


9547 
9595 
9643 
9689/9694 
9736/9741 
9782)9786 
9827/9832 
9872/9877 
9917/9921 
9961/9965 


3 


7427 
7505 
7582 
7657 
7731 


4, 


7435 
7513 
7589 
7664 
7738 


7810 
7882 
7952 
8021 
8089 
8156 
8222 
8287 
8351 
8414 


8476 
8537 
8597 
8657 
8716 
8774 
8831 
8887 
8943 
8998 


9053 
9106 
9159 
9212 
9263 
9315 
9365 
9415 
9465 
9513 


9562 
9609 
9657 
9703 
9750 
9795 
9841 
9886 
9930 
9974 


LOGARITHMS. 

5 6 7 8 9 
7443)7451 || 7459|7466)7474 
7520|7528 || 7536|7543|7551 
7597|7604 || 7612/7619)7627 
7672)|7679 || 7686/7694|7701 
7745)7752 || 7760|7767|7774 
7818|7825 || 7832|7839)7846 
7889|7896 || 7903/7910)\7917 
7959|}7966 || 7973/7980) 7987 
8028/8035 || 8041/8048/8055 
8096/8102 || 8109)8116)8122 
8162/8169 || 8176)8182)8189 
8228/8235 || 8241/8248|/8254 
8293/8299 || 8306/8312/8319 
8357/8363 || 8370/8376|8382 
8420/8426 || 8432)8439)8445 
8482/8488 |) 8494/8500/8506 
8543/8549 || 8555/8561/8567 
8603/8609 || 8615)8621|8627 
8663/8669 || 8675/8681/86386 
8722/8727 || 8733/8739|8745 
8779/8785 || 8791/8797/8802 
8837/8842 || 8848/8854/8859 
8893 8899 || 8904/8910)8915 
8949/8954 || 8960/8965/8971 
9004)9009 || 9015)9020)9025 
9058/9063 || 9069/9074/9079 
9112/9117 || 9122)9128/9133 
9165)9170 || 9175)9180/9186 
9217/9222 || 9227/9232/9238 
9269/9274 || 9279)9284/9289 
9320/9325 || 9330/9335/9340 
9370/9375 || 9380/9385/9390 
9420/9425 || 9430)9435/9440 
9469/9474 || 9479)/9484/9489 
9518/9523 |) 9528/9533/9538 
9566/9571 || 9576|9581/9586 
9614/9619 || 9624)/9628)9633 
9661/9666 || 9671}9675|9680 
9708/9713 || 9717|/9722/9727 
9754/9759 || 9763|9768/9773 
9800)9805 || 9809/9814/9818 
9845/9850 || 9854)9859)/9863 
9890/9894 || 9899}/9903)9908 
9934/9939 || 9943)/9948)/9952 
9978/9983 || 9987/9991/9996 


1 2534) 4. 6 (Gave Ss 


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bo bo bo bo bo bo bo bo bo bo 


— a et ee 


pt et ot 


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wmnodnhynwlrynmynnmnyn NWwwwwl wowwwow wowweo 


— at pt et 


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SSS oCcoorrF 


co eo CO CO 
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' 
tt pt pt 
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See ep See et tt et ht tt ee 
bo bo bo bo bo bo bo bO bO bo bo bo bo bo bo 


bo bo bo bo bo 


Oe ee a eG 


Proportional Parts. 


ee’ 


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bo bo bo bo bo 


bo bo bo bo bo bo bo bo bo bo bo bo bo bo bo 


ww We PP PP 


Cow WO 0 Co Go GO GO 


ww 0 0 Qo Oo OO GO OO 


bo bo bo Go CO 


bo bo bo bb bo bo bo bo bd bo 


[antl alls lls ls EPP PP PRR RR oon 


Oo GO GOD 


wwwww ew O93 0) OO 


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Co tg cl a ls ll Leal als eels oll sd PPP LP Pp He He HR OT Or Or Or Or Or Or Or Or Ot Or Or Or Otic & DD 2 2 


PORWOMOD Wsaivwsy 


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eee PP He Ee BR Or Or Or Ot Or Or Or 


CO ee 


PP PP LP 


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APPENDIX A 187 


The log of 0.483°°*? is found by multiplying the log by the exponent 
or 0.42 (9.6839—10) which is 4.067238—4.2. It is difficult to get 
the antilog of this directly, but the value of the logarithm is not 
changed if a number be added to the first part and subtracted from 
the second part to make this —1o. In this case add 5.8 to the first 
and substract 5.8 from the second, making the log of 0.4839” 
9.8672 —TIo. 

The antilog of this is 0.736 plus 0.0004 or 0.7364. 

That is, 0.483°°4? is equal to 0.7364. 


APPENDIX B 


Naperian Logarithms.—The natural, hyperbolic or naperian 
logarithm of a number can be found by multiplying the common log- 
arithm of the number by 2.3026 but the solution of problems involv- 
ing this log or the loge as it is written will be facilitated by the 
use of the following tables which read from 1 to 10 by increments 
of hundredths. 

For example, the loge of 4.36 is given directly as 1.4725. 

Characteristics and mantissas are not handled in this table in the 
same way as the common logs. But as the log of 43.6 is the same as 
the log. of 4.36 X10 this may be found by adding the logse of 4.36 
and to. In this case this is the sum of 1.4725 and 2.3026 or 3.7151. 
That is, the loge of 43.6 is 3.7151. . 

In the same way the loge of .436 is the same as the loge of (4.36 
divided by 10) or the loge of 4.36 minus the loge of 10. In this 
case it 1S 1.4725—2.3026 or —o.8301. That is the loge of 0.436 
is —0.8301, a negative number. } 


188 


PEAY DIX, B ; 189 


€ = 2.7182818 log e = 0.4342945 = M 
0 1 2 3 4 5 6 7 8 9 
1.0 {0.0000 |0.00995/0.01980,0. 02956/0. 03922\0. 04879|0. 05827/0. 06766|0. 07696/0. 08618 
1.1 /0.09531)0.1044 )0.1133 |0.1222 |0.1310 |0.1398 |0.1484 |0.1570 |0.1655 |0.1739 
1.2 |0.1823 |0.1906 |0.1988 |0.2070 |0.2151 |0.2231 |0.2311 |0. 2390 |0.2469 |0. 2546 
1.3 |0.2624 |0.2700 |0.2776 /0.2852 |0.2927 |0.3001 |0.3075 |0.3148 |0.3221 |0.3293 
1.4 |0.3365 |0.3436 |0.3507 |0.3577 {0.3646 |0.3716 |0.3784 |0.3853 » 720 |0.3988 
1.5 0.4055 |0.4121 |0.4187 |0.4253 |0.4318 |0.4382 |0.4447 |0.4511 {0.4574 |0.4637 
1.6 |0.4700 |0.4762 |0.4824 |0.4886 |0.4947 /0.5008 {0.5068 [0.5128 |0.5188 |0.5247 
1.7 |0.5306 |0.5365 |0.5423 |0.5481 /0.5539 /0.5596 |0.5653 |0.5710 |0.5766 |0.5822 
1.8 |0.5878 |0.5933 |0.5988 |0.6043 |0.6098 |0.6152 /0.6206 /0.6259 |0.6313 |0. 6366 
1.9 |0.6418 |0.6471 |0.6523 [0.6575 |0.6627 |0.6678 /0.6729 |0.6780 |0.6831 |0.6881 
2.0 |0.6931 |0.6981 |0.7031 |0.7080 |0.7129 |0.7178 |0.7227 |0.7275 |0.7324 |0.7372 
2.1 (0.7419 |0.7467 |0.7514 |0.7561 |0.7608 |0.7655 |0.7701 |0.7747 |0.7793 |0. 7839 
2.2 (0.7884 |0.7930 |0.7975 |0.8020 |0.8065 |0.8109 |0.8154 |0.8198 /0.8242 |0.8286 
2.3 |0.8329 |0.8372 |0.8416 |0.8459 |0.8502 |0.8544 |0.8587 |0.8629 /0.8671 |0.8713 
2.4 |0.8755 |0.8796 |0.8838 |0.8879 /0.8920 |0.8961 /0.9002 |0.9042 |0.9083 |0.9123 
2.5 |0.9163 |0.9203 |0.9243 |0.9282 |0.9322 |0.9361 |0.9400 |0.9439 |0.9478 |0.9517 
2.6 |0.9555 |0.9594 |0.9632 0.9670 |0.9708 |0.9746 |0.9783 |0.9821 |0.9858 |0.9895 
2.7 |0.9933 |0.9969 |1.0006 /1.0043 |1.0080 {1.0116 |1.0152 |1.0188 /1.0225 |1.0260 
2.8 |1.0296 {1.0332 |1.0367 |1.0403 |1.0438 |1.0473 |1.0508 [1.0543 |1.0578 |1.0613 
2.9 {1.0647 |1.0682 |1.0716 {1.0750 |1.0784 |1.0818 |1.0852 |1.0886 |1.0919 [1.0953 
3.0 {1.0986 |1.1019 (1.1053 |1.1086 )1.1119 |1.1151 {1.1184 |1.1217 {1.1249 /1.1282 
13.1 (1.1314 |1.1346 )1.1378 |1.1410 |1.1442 |1.1474 )1.1506 1.1537 |1.1569 /1.1600 
3.2 /1.1632 /1.1663 |1.1694 |1.1725 |1.1756 {1.1787 |1.1817 |1.1848 |1.1878 |1.1909 
3.3 {1.1939 |1.1969 |1.2000 {1.2030 |1.2060 {1.2090 [1.2119 |1.2149 {1.2179 |1.2208 
3.4 {1.2238 |1.2267 |1.2296 |1.2326 {1.2355 |1.2384 [1.2413 |1.2442 (1.2470 |1. 2499 
3.5 /1.2528 |1.2556 |1.2585 |1.2613 |1.2641 {1.2669 |1.2698 |1.2726 |1.2754 |1.2782 
3.6 [1.2809 |1.2837 |1.2865 |1.2892 |1.2920 {1.2947 |1.2975 |1.3002 {1.3029 /1.3056 
3.7 {1.3083 |1.3110 |1.3137 [1.3164 |1.3191 {1.3218 |1.3244 |1.3271 |1.3297 |1.3324 
3.8 {1.3350 |1.3376 |1.3403 |1.3429 |1.3455 |1.3481 |1.3507 |1.3533 |1.3558 |1.3584 
3.9 {1.3610 /1.3635 |1.3661 |1.3686 |1.3712 |1.3737 |1.3762 |1.3788 |1.3813 |1. 3838 
4.0 {1.3863 |1.3888 {1.3913 |1.3938 /1.3962 |1.3987 |1.4012 |1.4036 |1.4061 |1. 4085 
4.1 {1.4110 |1.4134 |1.4159 |1.4183 |1.4207 |1.4231 |1.4255 |1.4279 |1. 4303 |1.4327 
4.2 |1.4351 |1.4375 |1.4398 |1.4422 |1.4446 |1.4469 [1.4493 |1.4516 /1. 4540 |1.4563 
4.3 |1.4586 |1.4609 |1.4633 |1.4656 |1.4679 /1.4702 |1.4725 |1.4748 |1.4770 |1.4793 
4.4 |1.4816 |1.4839 |1.4861 |1.4884 |1.4907 |1.4929 |1.4951 |1.4974 |1.4996 |1.5019 
4.5 |1.5041 |1.5063 |1.5085 |1.5107 {1.5129 |1.5151 |1.5173 [1.5195 [1.5217 [1.5239 
4.6 |1.5261 |1.5282 |1.5304 |1.5326 |1.5347 |1.5369 |1.5390 |1.5412 |1.5433 |1.5454 
4,7 |1.5476 |1.5497 |1.5518 |1.5539 |1.5560 /1.5581 |1.5602 |1.5623 |1.5644 /1.5665 
4.8 |1.5686 |1.5707 |1.5728 |1.5748 |1.5769 {1.5790 {1.5810 1.5831 (1.5851 |1.5872 
4.9 1.5892 |1.5913 |1.5933 |1.5953 {1.5974 |1.5994 {1.6014 |1.6034 |1.6054 |1.6074 
5.0 {1.6094 [1.6114 |1.6134 |1.6154 |1.6174 {1.6194 |1.6214 |1.6233 (1.6253 |1.6273 
5.1 |1.6292 |1.6312 |1.6332 |1.6351 |1.6371 |1.6390 |1.6409 |1.6429 |1.6448 |1. 6467 
5.2 |1.6487 |1.6506 |1.6525 |1.6544 |1.6563 /1.6582 |1.6601 |1.6620 {1.6639 |1.6658 
5.3 |1.6677 |1.6696 |1.6715 |1.6734 |1.6752 |1.6771 |1.6790 |1.6808 |1.6827 {1.6845 
5.4 |1.6864 |1.6882 |1.6901 |1.6919 |1.6938 |1.6956 |1.6974 |1.6993 |1. 7011 |1.7029 
5.5 |1.7047 |1.7066 |1.7884 |1.7102 |1.7120 /1.7138 [1.7156 {1.7174 |1.7192 |1.7210 
5.6 (1.7228 |1.7246 |1.7263 |1.7281 |1.7299 |1.7317 [1.7334 [1.7352 |1.7370 |1. 7387 


190 AIR COMPRESSION AND TRANSMISSION 


NAPERIAN LOGARITHMS. 


eS et pe ——— a =" — pat pe 
ai hee! ve eee 8. . or ceils 


bO bO bO bo bo bo bo bo bo bo et _— — eee — eee —oe — aay 


bo bo bo bO bo bo bo DO bo bo bo bo bo bo bo bo bo eet oe Rt pet ——a— — ee — ee 


Nowy bobo bo bobo bo bo DNMp MMH MHP WH WH HYD Bee Se Eee 


Ddb MNMN Whwrp WO HWWHY HWwry 


COC CON AUR WHR © DON DOA WHH OO DON DOR WHH ©O CON QOR WHHL © SON 


D OOD WOOD OOO © MMM WMH MMH © NNN NNN NNN N Q2H 2QQ QAM @ aaa 


APPENDIX C 


HYGROMETRY'! 


Hygrometry is the measurement of the amount of water vapor in 
the atmosphere. There is always more or less water vapor in the 
atmosphere depending onits temperature and its degree of saturation. 
The study of hygrometry is of increasing importance. It has been 
found by experience that the moisture in air has a marked effect on 
many industrial processes, suchas the spinning of cotton, the smelting 
of iron in blast furnaces and the ventilation of factories and other 
buildings. It isalso necessary to know the amount of moisture pres- 
ent in all measurement of air or gases and in tests of machin- 
ery for handling the same. 

According to Dalton’s law, when a mixture of two gases fills a space 
of, say, 1 cu. ft., the pressure in the space is the sum of the two pres- 
sures that would be produced by a cubic foot of each of the gases alone 
at the same temperature. In the same manner a mixture of air and 
vapor hasa pressure which is the sum of the pressure of an equal vol- 
ume of dry air, and of vapor alone, each at the given temperature 
of the mixture. Air and vapor occur in mixtures varying from prac- 
tically dry air to a state of saturation such that any addition to the 
mixture of vapor at the same temperature causes a portion to con- 
dense. To every temperature there corresponds a certain water- 
vapor pressure or partial pressure which may be found in steam 
tables such as ‘“‘ Marks and Davis” or ‘‘ Peabody’s.”’ 

Air in actual practice rarely contains vapor with 1oo per cent. sat- 
uration and the weight of water vapor present is less than the maxi- 
mum for that temperature of air. The air is then said to be only 
partially saturated, and the degree of saturation is expressed by the 
ratio of the weight of water vapor actually contained in a given space 
to the maximum weight that the space can contain under the condi- 
tions of absolute pressures and temperatures existing at that time. 
This ratio is known as the “Relative Humidity.” 

Absolute Humidity.—The absolute humidity is the weight of water 


1 Christie’s and Kowalke’s Steam and Gas Engineering Laboratory Notes. 
191 


192 AIR COMPRESSION AND TRANSMISSION 


vapor that 1 cu. ft. actually contains under the given pressure and 
temperature conditions. 

Relative Humidity.—Relative humidity is usually determined by 
means of psychrometers or wet -and dry-bulb thermometers. These 
consist of two thermometers fastened to a frame and placed in a 
current of air. The bulb of one thermometer is kept covered with 
cotton wick and is kept thoroughly wet with water at room tempera- 
ture. Ifthe airis not saturated, evaporation will take place from the 
wet bulb and its temperature will be lowered by the abstraction of 
the latent heat of the water. This lowering of the temperature 
has been found to be a measure of the relative humidity. 

Psychrometers.—Psychrometers are made in two types, stationary 
and sling. In the sling psychrometer the wick is moistened and the 
whole frame whirled around by a handle for 15 or 20 seconds. The 
wet bulb thermometer is read immediately after stopping. By the 
use of Chart A of the accompanying diagram the relative humidity 
may be obtained from the readings of the two thermometers. For 
example, if the dry-bulb thermometer shows a reading of 72° and the 
wet bulb 61°, or a difference of 11°, the relative humidity is 52.6 per 
cent., if the atmospheric pressure is 30 in. of mercury.. Ifthe atmos- 
pheric pressure is 28 in. the chart reading should be increased 2X 1/100 
of 52.6 or 1.05, making the corrected relative humidity 53.65 per cent. 

In making accurate measurements of air it is necessary to deter- 
mine carefully the weight of moisture present in the atmosphere and 
the volume occupied by this vapor. In order to do this, reference is 
made to data regarding the density of vapor at various temperatures 
and pressures. This information is given in most steam tables and 
the following figures have been taken from Marks and Davis tables. 
The density or weight of the vapor per cubic foot isshown graphically 
in Chart C of the large diagram as the line marked 100 per cent. 
This same chart also shows weights of the vapor to be used in calcu- 
lations with air of various humidities, as shown in the example fol- 
lowing’ the tables. 


— ee ee 


“= Le ee 
SS te — < “4 : ‘ Te 


Bs we 2 7 = 


een ae 


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tf 


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+32 § ape IN ie ner ; no eM ie Ut Oh OSes ey Sis se eS 
a & aes im 2% > ‘ .* 
a" o> fa: x 4 Er “t-i-* 4 S > ; ' 
“Fy dee he ee res ~ ¥ a - % - oe a + Anes — 
ie. waice OL , ‘ ; 4 S . eH ' r ' : i ¥ 
*% Oe ela She Se PT : _ ee : bigs ek Poe be) ee 
aK % al i er Os: es CSAP Ly anes Se SRS CRN ete er Cone = 
= fs . ‘ i , t I oo 
oer BEN, t en — SOE ee a Sy ere 1 
SESS 2 Soto eae 
a ss = = “ e = \ ~ aa a 
SSFP a) SIN 1) sear WES COA) SEY SN NNER a Saeco (OOS ESET eS Re Boe a Ee 


‘ 


Temperature of Air or Dry Bulb Thermometer, Degrees Fahrenheit 
coy o fo} o So 


100 F = = = =: 

= 

£ 

5 | 
= >! pe 

= 2 

1 

ao 

= 3 
mod 

2 1 4 
= 

ry 5 
S Fe 

” 

= 7 
3 

E a 
5 aS 
&S = v 
eer 0 = > § 
Pe =< 

xk = 10-5 
of = 
Ee ~v nea 
oS — 
ae ~ > 
3 Fy es 
+ Dil 

5 eze : 88 
EES z 
pee bans = = 4 5 
o gt =< 2 
+fo = cy 
© Sts = Ibi 
ae =. © 
aS Pte les 
£ 3g 50 Ss 
oo? ~ 7 3 
eg = = 
Ss - 
-o¢ 18 @ 
PA See A a 
Les 19 
mo] rT) 
E2555 a: 
282" 71 Ps "s 
rae y~ (er 
oo = o 
= 230 
ees 1 = al > 
Ena So ~ 4 8 
-—-> - 3 

ot 3 = 25 = 
«3530 = = 
Ew ~ 75 
<P EY A 27 © 
£3 fe S 2 £ 
Se : H- a a= 
$3 “SI SvAwy, 30 
e 00 A-X.. 31 
35 s /\ A 32 
aE a 

Po aa 33 
at = 134 

(=) = 

eee tes eae 35 
&z> ~F-L17 36 
Ea ld e (oa 3 
ow 38 
=k f 

S 39 

Yo Oo 

Lo ss 40 
ce 4l 
ov ;’ 42 
vo 

ge Bi 
-=3 0 

CHART A 


rey re ° 2 
is) Ss Ss So So is 
105 of, 6 
is F : | | TTT y | ourune | | [ Lop 
; | OUTLINE OF CALCULATIONS ; 
Specific Heat of Air ot 14.5 Lb. per Sq. Ih: + Weight of Water Vapor in Pounds} per Cubic!Foot+ = 2 2 ae. ae + i al Teoh Zz 
P 7B. TU; perPound PSb-oe Ieatcvintea from Valves givenin Marke a Dowie Steam Tables i i : 4 das roe ees ieee al a, hu 
oO w uw ‘a = : . i e A 
Bal ae | lies i Ss S S S | S 3 By Sturated vapor pressure in lbs. per sq. in.(Marks &Davis Stearn Tables used) | we 6 lo or AS? ASD Ay AS 40% 0 fo coe or 
S = B ° = Ss S Ss Ss Ss xP r Cent Hurniality | | | | | hae 00045 
100 = f <Teynperature of Air in degrees Fahrenheit fel alee 12 = 
W= Weight of one cu.ft: of a mixture of alrand water vapor at t-termp, press. ana xR humidity 7-s 2g 
3 = Wa- Weight of one cu ft of dry airata pressure of ('40-xp,) las. per Sq-171- t$—}— S = 48 
S We = Correction tobe added to ve for pressure above 40 /bs. per sq.in. | = + Lit ce 
= ely 2 Jest | Wy Weight of water vapor contained in ane cu.ft of saturated air | | z 
ion S\3 eS S\S Ss S88 = S =Specific heat ofa mixture of air and water vapor(B.tu//b. ) a =f Le 
S Man Vink Wa a Sa Smecitic heat of dry alr=Q242+0.000009 t: Harvey M Davis, FarnsASME, Vol50,e 750/908) 3 r B 0040 & 
Sw-Spyecitic heat of water vapor-04423+0 00018 WilisH Carrier Jour ASME, Vol33 2328/91) Ss = fu 
= R-SRIS | | fe 3 it £ 
T 459.644 —| | te 2 a 
ifeel Fo Dry Alc: + 25 - => Se] a Ea 
1 Pala |RT 5 Wee pei W= gt Worx My pe 2 — S ie 33 
ge GED AOA oud) 3 8 0.0035 ++ < 
90 F5S5|4596+7 ilies! $ [2 = 
+ =i =\5 ~—t oO a 
nee 448 14.0) x £ o£. 
(OB S5\45S9.6 +4) - | a haa a ion ib s 3 
L aoa Wo + We) Ser) + Wy Sw) a S + S fe S s 
= Way? WeFX 3 Ss 
ie i+ ‘cal at VE PX My, Ss > 
p + = 0.0030 i 2 
£ 85 3 a alae 2 
& 3 S ¢ > £ 
4 i 5 S = co 4° ae 
8 2 4 | o wo oe 
2 ey a 1 cS 8 tg = ~ 
+= aH S 
S. | Ta r Cor re + x r3 = $ 
a I SSS AOS ee Se c 0.0025 S18 lst! 3 
fe, | LT LY ree |S Tr T || oe s+ rt ‘ 
= AO = E <> OF ue 
S ie | | =a x | ‘a & | AS esc ty Son a > 8 
2 | |_| les we ae i 1 fe af e Ss 
g = + =H e Fe ii 164 te oe. Re 
= & Mo & 
Bas a ee Sat ret = a 7 ao020 "4 ‘$ & 
: a | : Hts bs Hee EHH Z 
£ a Sales = 2 Sy rete re = i 
& + | “ 2 ae jee ee 13 RG #3 : sleet & 
RG S ~ 
Rs S| tt Ie Lt + o & a = 5 
a {its SLE re mS I &A8 g 
70 eS rast i Nes ee 2 
cs Heer 1 « AS Lag 
is Teo S | es | | ae ~Oy eS | 4 
: at = Tl Ss 7 hey * rae Se a e 
S 7 y & %S 
pee anee fooes Betts HEE 
) SEE ee eet Cees . s we S 
65 x 8 | - Ly | =e elms} _ & 
5S ; x S) 
aE Ss 4—S ie ey 
[ : t +. ttt eet { : 
Ct ae POPE ae a et et 5 ; 
= = T ir + + Al | as roe | 3 oa < 
= 2 | ©) = 
| EEC | aS 7 Tt ; 
60 ] leaatssye | ie a CI > = | : L S 
« < 
a t * es " = 
Error in Specific Heat Curves: ; = oe es ee 
4100 °F, 100% hurnidity anid pressure of #0 /é or- a at S = e: TI] gs ima iS | 
/5.0 |b per sg.in. these vizlves are in errorgof orre %. Bil ie Js iS | [ ze ' 
At 60°F, 00% humidity arid pressure of 140 /b or + ie [ [ | i © H 
18.0 Ib. per sg.in. these values are in errorg ofone % = tn : +—|—+ “tt Ve | et | 2 
55 S w ne % . 
CHART B CHART C zg : GHART.D * CHART E 8 CHART F g CHART 6 
sg s 3 s 


Humipiry, WEIGHTS PER CuBIC Foor AND SPECIFIC HEATS oF 


W,. C. RowsE, Mapison, Wis. 


1 


TIXTURES OF AIR AND WATER VAPOR 


= 105 


95 


90 


ao 
a 


s 
erature of Air, Degrees Fahrenhe 


~~ 


Temp 


65 


60 


55 


APPENDIX, C 193 


Tables showing the temperature, pressure, specific volume and density of 
steam or water vapor from 32° to 219° F., condensed from Marks’ and Davis’ 
Steam Tables by permission of the publishers, Longmans, Green & Co. 


pO Specific vol., Density, 
Temperature : 
cea cubic feet pounds per 
Pounds per Inches per pound cubic foot 
square inch mercury 

32 0.0886 0.1804 3,204 ; O.000304 
2 2A Be 0.0922 0.1878 ; pL Om ee 19 OC. 0003T0 
34 0.0960 0.1955 3,052 0.000328 
35 0.0999 0. 2034 2,938 ©.000340 
36 0.1040 Oar rn? 2,829 0.000353 
27, 0.1081 - 0.2202 ayn 0.000367 
38 O.1T25 O.2290 2,626 0.000381 
30 Oni170 0.2382 2,530 0.000395 
40 OLT227, OV2477 2,438 0.000410 
41 0.1265 0.2575 pe Ke. 0.000425 
42 0.1315 0.2677 2,266 0.000441 
43 0.1366 On 2782 2,185 0.000458 
44 0.1420 0. 2890 2,107 0.000475 
45 OfT 47s ©. 3002 2,033 0.000492 
46 OLDE 32 0.3118 1,961 ©.000510 
47 0.1591 0.3238 1,892 ©.000529 
48 0.1051 0.3363 1,826 0.000548 
49 0.1715 0.3492 1,763 0.000567 
50 0.1780 0.3625 1,702 0.000587 
51 0.1848 0.3762 1,643 0.000608 
52 O.1Q17 0.3903 1,586 0.000630 
53 0.1989 ©. 4049 1,532 0.000653 
54 0.2063 0.4201 1,480 0.000676 
55 0.2140 0.4357 1,430 2.000700 
56 0.2219 0.4518 1,381 0.000724 
i) 0.2301 0.4684 335 0.000749 
58 0.2385 0.4856 1,291 0.000775 
59 Os 2472 0.5034 1,249 0.000801 


13 


194 AIR COMPRESSION AND TRANSMISSION 


Pate Specific vol., Density, 
Temperature : 
re tometer cubic feet pounds per 
Pounds per Inches per pound cubic foot 
square inch mercury 
60 0.2562 0.522 1,208 0.000828 
61 O2054 0.541 1,168 0.000856 
62 0.2749 0.560 1,130 0.000885 
63 0.2847 0.580 1,093 ©.O000Q15 
64 ©. 29049 0.601 1,058 0.000946 
65 0.3054 oO. 022 1,024 0.000977 
66 0.3161 0.644 ggI ©.O001009 
67 0. 3272 0.667 959 ©.001043 
68 0.3386 0.690 928 O.001077 
69 0.3504 On7iA 899 O.OO111I2 
70 0.3626 0.739 871 0.001148 
71 Bgat 0.764 843 0.001186 
7 0. 3880 0.790 817 O.001224 
Fes O.4012 0.817 792 0.001263 
74 0.4148 0.845 767 ©.001304 
ws 0.4288 0.873 743 0.001346 
76 0.4432 0.903 720 0.001389 
ot 0.4581 0.03% 698 0.001433 
78 0.4735 0.964 677 0.001477 
79 0.4893 0.996 657 0.001523 
80 0.505 I.029 636.8 ©.001570 
8I O15 22 ip O0s O17 5 0.001619 
82 0.539 1.098 598.7 0.001670 
83 OL587 TLi134 580.5 O.001723 
84 OVS75 ie 7a 562.9 O20017 77 
85 0.594 I. 209 545.9 0.001832 
86 O01? 1.248 526:.,5 0.001889 
87 0.633 1.289 Bay ©.001947 
88 0.654 1 B31 498.4 ©.002007 
89 C2075 Daas 483.6 0.002068 
go 0.696 TgA17 469.3 O.002131 
QI °.718 1.462 455-5 0.002195 
92 OL7AT 1.508 442.2 0.002261 
93 0.765 T2556 429.4 0.002320 
04 0.789 1.605 417.0 0.002398 


APPENDIX C 195 
Pressure . : 
Specific vol., Density, 
Temperature 2 
Fahrenheit cubic feet pounds per 
Pounds per Inches per pound cubic foot 
square inch mercury 

95 0.813 1.655 405.0 0.002469 
96 0.838 1.706 303-4 0.002542 
97 0.864 Le 7S0 20242 0.002617 
98 0.891 1,813 27 LMA 0.002693 
99 0.918 1.869 360.9 0.002771 
100 0.946 1.926 350.8 0.002851 
IOI 0.975 1.985 GAO 0.002933 
102 1.005 2.045 ZZ 1s 0.003017 
103 TROZS Siew B2242 0.003104 
104 1.066 py hs Sy iew 0.003192 
105 1.098 22230 BO4m 7) 0.003282 
106 Pers 26303 2960.4 0.003374 
107 1.165 gus72 288.3 0.003469 
108 I.199 e443 280.5 0.003565 
109 re235 2eSLs 27250 0.003664 
IIO Lear 2.589 265.5 0.003766 
III 1.308 2.665 2522 0.003871 
Lis 1.346 2740 oa Ret) 0.003978 
by he 14386 2,822 244.7 0.004087 
II4 1.426 2.904 238.2 0.004198 
II5 1.467 2.987 2270 0.004312 
116 I.509 3.072 22573 0.004429 
ray 14552 Be ror 219.9 0.004548 
118 1.597 2-252 ZrAe TL 0.004671 
I1g 1.642 3.344 208.5 0.004796 
120 1.689 3.438 203.1 0.004924 
I2E 1.736 aoe 197.9 0.005054 
22 e705 3.635 192.8 0.005187 
123 1.835 By Ta 187.9. 0.005323 
124 1.886 3.841 183.1 0.005462 
125 1.938 3.948 178.4 0.005605 
126 1.992 4.057 17300 0.005751 
127 2.047 4.168 169.6 ©.005900 
128 21O 4.282 16553 0.006052 
129 2.160 4.399 POT SI 0.006207 


196 AIR COMPRESSION AND TRANSMISSION 


Pena Specific vol., Density, 
Temperature ; 
ea Nes cubic feet pounds per 
Pounds per Inches per pound cubic foot 
square inch mercury 

130 2270 AasZ TS 7% 0.00637 
131 221276 4.64 a5 a2 0.00653 
132 2.340 4.76 149.4 0.00669 
133 2.403 4.89 145.8 0.00686 
134 2.467 5.02 142.2 0.00703 
135 2.533 5.16 13027 0.00721 
136 2.600 5.29 135.4 0.00739 
£37 2.669 5-43 Poza 0.00757 
138 R740 5.58 128.9 0.00776 
139 2.812 ew ke 125.5 0.00795 
140 2.885 5.88 122.3 0.00814 
141 2.960 6.03 119.9 0.00834 
142 Bk027 6.18 fi721 0.00854 
143 Seni as 6.34 114.3 0.00875 
144 BOs OF 55 IIr.6 0.00896 
145 Bee 77 6.67 109.0 0.00918 
146 Be ZOT 6.84 106.5 0.00940 
147 3.446 702 104.0 0.00962 
148 23:3532 7220 101.6 0.00985 
149 3,023 7.38 99.2 0.01008 
150 3.714 2A57 96.9 O,O1L0s2 
I51 3.809 7 270 04.7 0.01056 
152 3.902 7.95 92.6 0.01080 
153 3-999 8.14 9005 = O.O1I05 
154 4.098 8.34 88.4 O;OEI3I 
155 4.199 Gas5 86.4 O.OII57 
156 4.303 8.76 84.5 0.01184 
r57 4.408 8.98 82.6 O.OI2I1 
158 4.515 Q.20 80.7 0.01239 
159 4.625 Q.42 78.9 0.01267 
160 4.737 9.65 7702 0.01296 
161 4.851 9.88 7S uS 0.01325 
162 4.967 10.12 fc tates O.01355 
163 5.086 10. 36 FORD 0.01386 
164 5.208 10.61 70.6 O.OI417 


Temperature 
Fahrenheit 


APPENDIX,C 197 
Pressure : ; 
Specific vol., Density, 
cubic feet pounds per 
Pounds per Inches per pound cubic foot 
square inch mercury 

5-333 10.86 69.1 0.01448 
5-460 LLere 67.6 0.01480 
5-589 T1245 66.1 O.OI1513 
cana LI.65 64.7 0.01546 
5.855 II.Q2 63.3 0.01580 
5.992 12520 62.0 0.01614 
OL 135 12.48 60:7 0.01649 
09273 iy | 59.4 0.01685 
6.417 13.07 58.1 O-On7 21 
6.564 E3437 56.9 0.01758 
6.714 E3307 Bey 0.017096 
6.867 13.98 54.5 0.01834 
i; O23 14.30 53-4 0.01873 
Teth2 14.62 cos 0.01912 
7-344 14.95 51.2 0.01953 
fhe wt 15.29 5O. 15 0.01994 
7.68 15403 49.12 0.02036 
TEs 15.098 48.12 0.02078 
8.02 16.34 47.14 O-O2T2T 
8.20 £570 46.18 0.02165 
8.38 17.07 Aseas 0.02210 
8.57 17.45 44.34 0.02255 
8.76 L7os 43.45 0.02301 
nOE05 10.22 42.59 0.02348 
9.14 18.61 ALTA 0.02396 
9.34 19.02 40.91 0.02444 
9.54 19.43 40.10 0.02493 
9.74 19.83 39.31 0.02544 
9.95 20.27 38.54 0.02505 
TOL 7 20n7t 37.78 0.02647 
10.39 Zia Ds 37.04 0.02700 
10.61 21.60 3632 0.02753 
10.83 22505 acyo? 0.02807 
II.06 PUNE FAnOS 0.02863 
71.20 22.99 34.26 0.02919 


198 AIR COMPRESSION AND TRANSMISSION 


Eres Specific vol., Density, 
Temperature ’ 

Pahreoheit cubic feet pounds per 
Pounds per Inches per pound cubic foot 
square inch mercury 

200 tT ah 2 225A 7 33.60 0.02976 
201 11.76 23.95 32.96 0.03034 
202 12.0% 24.45 3 2e a3 0.03093 
203 12520 24.96 2272 0.03153 
204 12555 25.48 Gratz 0.03214 
205 T2477 26.00 20,53 0.03276 
206 13203 20.53 29.95 0.03339 
207 i3430 27.08 29.39 0.03402 
208 Tig 554 2763 28.85 0.03466 
209 13505 28.19 Baa 0.03531 
210 T4512 28.76 27.80 0.03597 
Bit 14.41 20.33 27-20 0.03664 
Bee I4.70 29.92 26.79 Of02732 
213 14.99 202 26.30 0.03802 
214 15.29 ar 1s 25202 0.03873 


Partial Pressures.—Suppose, for example, the barometers read 
29.214 in. of mercury at a temperature of 78° F. Chart F of the 
diagram shows that at this temperature 1 in. of mercury corresponds 
to a pressure of 0.4889 Ib. per square inch. That is, the barometer 
reading of 29.214 in. of mercury corresponds to an absolute pressure 
of 14.2827 lb. per square inch. If the air is saturated with moisture 
at 78° F., the pressure exerted by this vapor is, as shown from the 
tables of Marks and Davis, 0.4735 lb. persquareinch. The pressure 
of the dry air present would then be 14.2827—0.4735 or 13.8092 
lb. per square inch. 

Suppose the psychrometer shows a relative humidity of 40 per 
cent. As the vapor pressures are proportional to the absolute 
weights, the pressure exerted by the moisture in the air will be 
40 per cent. of 0.4735 or 0.1894 lb. per square inch. In this case 
the pressure due to the dry air present will be 

14.2827 —0.1894 or 14.0933 lb. per square inch. 

If it is necessary to find the weight of a cubic foot of this moist 

air, this can be found by adding the weight of the cubic foot of dry 


ALPEN DIX'G VES}) 


air at its pressure and temperature to the weight of the vapor 
present. 

The weight of vapor present is found by multiplying the weight 
of a cubic foot of vapor at the given temperature by the relative 
humidity. The tables show that 78° F., the weight of a cubic foot 
of vapor, is 0.001477. The weight of the vapor present in the 
example is 40 per cent. X0.001477 or 0.000501 lb. 

The weight of dry air present is found from the formula 


BiVa 144 X 14.0933 
Sh Sys == - =0.070 
53-311 53-3(460+78) Nie 


The weight per cubic foot of the air and its accompanying vapor 
is 


== le 


0.000591 +0.070773 =0.071364. 


This calculation can be made quite simply by referring to the 
various charts of the large diagram. By referring to Chart D it 
will be seen that the weight of air at 4o per cent. relative humidity 
and 78° F. is .o6992 lb. per cubic foot if the pressure of the atmosphere 
is 14 lb. per square inch. In the example given the pressure is 
14.2827 lb. per square inch. By referring to Chart E it will be 
seen that for the pressure of 14.2827 and temperature of 78° F. a 
correction of 0.00144 should be added making the weight per cubic 
foot of this mixture 


0.06992-+0.00144 or .07136 lb. per cubic foot. 


When it is desired to measure air with a Thomas electric meter, 
the mean specific heat of the mixture of air and water vapor must 
be known. W. H. Carrier in his paper ‘‘ Rational Psychrometric 
Formule,” Journal A. S. M. E., Nov., 1911, gives the following 
values which represent the results of the more recent investigations 
on the specific heat of air and water vapor. Instantaneous specific 
heat of air 


C pa =0.24112-+0.0000001 


where ¢ is the temperature in degrees Fahrenheit; and the instan- 
taneous specific heat of water vapor as approximately 


C ps =0.4423 +0.0001 8 


where ¢ is the temperature in degrees Fahrenheit. 
Applying these formule to the example given with temperature 
of 78, C pa IS 0.241822 and C ps 1S 0.45634. 


200 AIR COMPRESSION AND TRANSMISSION 


The mean specific heat can then be found by multiplying the 
weight of each substance in the mixture by its specific heat, adding 
the products, and dividing the sum by the weight of the mixture. 
Thus 


For the air, 0.070773 X0.241822 =0.017114 
For the moisture, 0.000591 X0.45634 =0.000270 


0.017384 


Mean specific heat is 0.017384 -+0.071364 or 0.2436. 

The mean specific heat may also be obtained by referring to 
Chart B of the large diagram. This shows that for the given temp- 
erature of 78° F. and a relative humidity of 40 per cent. the mean 
specific heat may be taken as 0.2435. 

The above principles are applied commercially in testing steam 
condensers. An accurate thermometer is placed in the suction 
to the dry air pump and a mercury column attached to the same. 
In a condenser the conditions are such that the mixture is always 
saturated. Hence the pressure due to water vapor passing to the 
air pump will equal that due to its temperature as given in the 
steam tables. Then the difference between this pressure and that 
shown by the mercury column will equal the pressure due to the 
dry air in the mixture. If the volumetric efficiency of the air 
pump is known, the amount of air pumped can be computed, and 
this gives a means of readily checking the condensing equipment 
- for air leakage. 

The large diagram containing Charts A, B, C, D, E, F and G 
was prepared by W. C. Rowse, Instructor in the Steam and Gas 
Engineering Department of the University of Wisconsin. 


INDEX 


Absolute humidity, 191 
temperature, 5 
zero, 6 
Action of piston compressor, 70 
Actual card of piston compressor, 78 
compression, 75 
Advantage of isothermal compressor, 
25 
of multi-stage compressor, 90 
AIT UE 
at low pressures, 38, 68 
at pressures below the atmosphere, 
26,—68 
composition, 1 
characteristics, 1-4 
and energy equations, 
10-17 
compressor cards, 75 
discharge valve, 102 
density at various pressures, 174 
dry, 4 
for cupolas, 39 
for forges, 39 
for ventilation, 39, 40 
free, 2 
humidity, 2-6 
internal energy, 6, 7, 16 
in water, 29 
inlet valve, ror 
measurement, 160-171 
pump, Edwards, 31 
U.S. Navy, 30 
supply for various buildings and 
rooms, 40 
Allis Chalmers fan, 65 
Altitude effect, 140-144 
Anemometers, 43 
Apparatus for measuring large quan- 
tities of air, 166 
Apparent specific heat, 8 
volumetric efficiency, 77 
Area of inlet valves, 100 
of discharge valves, 1or 
of fan blast, 43, 62 


Arrangements for coupling  turbo- 
blowers, 125 

Arthur compressor, 132 

Arthur, Thomas, 132 

Automatic valves, 100 

Available power, 179 

Axial discharge fan, 41 


thrust, balancing, 121 


Balancing axial thrust, 121 
Rateau impellers, 122 
by balancing piston, 123 
by counter position, 121 
by diminishing back area, 122 
Baloche and Krahnass compressor, 
eWay Bed 
Belt regulator, 105, 107 
Blast area, fans, 43, 62 
Blower capacities, 50 
cross section, 50 
definitions, 42 
efficiency, 81 
losses, 81 
mixing, 127 
Parsons, 114 
pressures, 50 
Rateau, 114 
Blowing engine, 41 
Blowers, 41 
Boyle’s law, 10 
Brake horse-power for fans, 58, 60, 66 
Brauer’s method of constructing ex- 
ponential curves, 19 
Brown, Boveri and Co. turbo-com- 
pressor, 117 
British thermal unit, 6 
Buildings, air required, 40 


Calculated and actual horse-power 
required for single stage com- 
pression, 74 

Capacity of blowers, 50 

of fans, 42 
of intercoolers, 93, 94 


201 


202 


Capacity of receivers, 160 
Card of piston compressor, actual, 78 
ideal, 77 
Cards, combined two-stage, 147 
clearance unloader, 112 
from air compressors, 70, 75 
showing adiabatic and isother- 
mal compression, 73 
Carrier, W. H., 199 
Centrifugal fans, 38-65 
Channing, J. Parke, 144 
Characteristic and energy equations 
fOtedit, 20-87 
equation for perfect gas, 
IO 
Characteristics of air, 1-4 
Christie, A. G., 101 
Classification of fans and blowers, 41 
of valves, 98 
Clayton governor, I09 
Cleaning valves, 182 
Clearance effect, 70, 71, 96, 97, 990 
methods of reducing, 71 
unloader, 110, 112 
> Cards 142 
Coefficient of contraction, 43 
of efflux, 43, 56 
of velocity, 43 
Combined cards, two-stage compressor, 
147 
governor and regulator, 109 
Common logarithms, 184-186 
Comparative effect of altitude on out- 
put, 143 
Compensator, hydraulic, 83 
lever, 83, 84 
weight, 83 
Composition of air, 1 
Compressed air explosions, 182 
Compression, actual, 75 
isothermal, 25 
line, 73 
wet and dry, 74 
exponential, 23 
Compressor, direct-acting steam, 82 
low pressure, 38 
tests, 144, 158 
Computation of internal or intrinsic 
energy. 16 


INDEX 


Concentration of liquors, 34 
Condenser pumps, 27 
Cone wheel fans, 65, 66 
Constants for pipe formule, 174,175 
Construction of equilateral hyperbola, 
18, 19 

of exponential curves, 19 

of isothermal curves, 18 
Contraction, coefficient of, 43 
Cooling capacity, 93 

devices, 117 

surface, 93 

turbo-compressors, II5 
Cost of Taylor compressor at Ains-_ 

worth, B. C., 136 

Coupling compressors, 124 
Cross-section, standard blower, 50 

piston compressor, 69 
Cupolas, air required, 39 
Cutler-Hammer Co., 161 
Cylinder efficiency, 80 


D’Auria system of energy compensa- 
tion, 83 
Dalton’s law, 191 
Davis, G. J., 164 
Definitions, fundamental, 5—9 
for fans and blowers, 42 
Density of air for various pressures, 174 
of water vapor, 193-108 
Description of fans, 58 
Design of fans, 58, 67 
of turbo-compressors, 113 
Details of piston air compressors, 98- 
110 
Developed section of Parsons blades,115 
Devices, cooling, 117 
Diagram, three stage piston compres- 
sor, 116 
turbo-compressor, 116 
Diagrammatic sketch of Thomas elec- 
tric meter, 169 
Diagrams, graphical, 18-25 
Difference between isothermal and 
adiabatic compression, 22 
Direct acting steam compressor, 82 
Disc fan, 58 
Discharge from a fan, 57, 59, 66 
valve, 102 


INDEX 


Discharge, area, or 
Draft measurement, 43 
Dresser coupler, 172 
Dry air, 4 
pump, 27 
Duplex compressor, 86 
cross compound steam, two-stage 
air compressor, 88 
belt driven compressor, 87 
steam driven compressor, 87 
Durleys Ref. 266 


Economic efficiency, 81 
Edwards air pump, 31 
Effect of altitude, 140-144 
of clearance, 70, 71, 96, 97 
of changing discharge pressure, 99 
of early closing of inlet valve, 73 
of pressure on temperature, 4 
Effects of heat, 6 
of outlet on capacity, 55 


Effects of pressure on  tempera- 
ture A 
Efficiency, apparent volumetric, 77 
blower, 81 


cylinder, 80 

economic, 81 

of compression, 80 

of fans, 45 

of Taylor compressor, 134 
Efficiencies, 77-82 

true volumetric, 80 
Efflux, coefficient of, 43, 56 
Electric meter, diagram, 169 
Energy, 5 

compensation, 82-88 

in air, 6 
Engineering Magazine, 113 
Equalizing steam pressure and air 

resistance, 82 

Equilateral hyperbola, 18, 19 
External energy changes, 6 
Expansion of casing, 118 
Explosions, compressed air, 182 
Exponential compression, 23 

curve construction, 19 


Fan, blast or steel plate, 60 
capacity, 42 


203 


Fan, centrifugal, 38-65 
cone wheel, 65-66 
definitions, 42 
design, 58-67 
description, 58 
discharge, 58, 59, 66 
efficiency, 45 
losses, 45 
mechanics of, 52 
pressure, 42 
proportions, 41, 61 
radial wheel, 58 
speed, 62, 67 
Fans, axial, 41 
classification of, 41 
or blowers, 41 
Flow of gas through an orifice, 45, 46 
Forges, air required, 39 
Forms of poppet valves, ror 
Free air, 2 
discharge, 42 
Friction effect of elbows, 61, 176 
Frigells |5.P:.120 
Frizell’s compressor, 129 
Fundamental definitions, 5-9 


Gases in air, I 
Governor and regulator combined, 109 
Clayton, 109 
for electric driven compressors, 
107 
Nordberg, 109, 110 
Grains, vapor per cu. ft. saturated air, 2 
Graphical construction of exponential 
curve, 18, 19 
of isothermal curve, 18, 19 
diagrams, 18-25 
method of determining 
head, 165 


mean 


Halsey, F. A., 142 
Hammon coupler, 172, 173 
Heat. 5 
added or taken away for iso- 
thermal change, 21 
for exponential change, 21 
etrects.6 
taken away during compression, 
22 


204 


Hero’s device for opening temple doors, 
VII 
fountain, VII 
Horse-power, brake for fans, 58, 60, 66 
single-stage compression, 74 
Horizontal-vertical arrangement of 
cylinders, 86 
Housing for fans, 42-63 
Humidity, absolute, ror 
OL aipmtone 1s & 
Hydraulic air compression, 129-139 
air pump, 26 . 
compensator, 83° * 
compression losses, 138 
compressor, Arthur’s, 132 


Baloche and Krahnass, 131, 132. 


Taylor’s, 133-137 
Hygrometry, 191 


Ideal card, piston compressor, 77 
Impellers, rotary blowers, 49, 50 
Improved cooling, turbo-compressors, 
118 
Indicator card piston compressor, 70 
cards, condenser pumps, 30 
Industrial uses vacuum, 32 
Ingersoll Rand Co., 103, 111, 112 
compressor, 147 
Inlet connection, 183 
for blowing fan, 61 
for exhaust fan, 61 
valve, Ior 
area, 100 
setting, Ior 
Intercoolers, 90 
capacity, 93 
Nordberg, 92 
pressure, 93 
surface required, 93 
types, 92 
tubes, 92 
with separator, 92 
Internal energy changes, 6 
or intrinsic energy of air, 7 
computation of, 16 


acver. (5 Heerco 
Jaeger’s turbo-blower, 119 
patent impeller, 120 


INDEX 


Kennedy blowing engine valve, 105 
Kowalke, O. L., 191 
Krahnass, A., 131 


Labyrinth bushing, 120 
Law, Boyle’s, 10 
of Charles, ro 
Leakage past turbo-stages, 120 
Lecture by H. deB. Parsons on fans, 
41-68 
Lever compensation, 83, 84 
Leyner air reheater, 177 
Liquors, concentration of, 34 
Logarithms, common, 184-186 
Naperian, 188-190 
Loss of capacity due to clearance, 79 
of head due to friction in ducts, 47 
Losses of blower, 81 
of hydraulic compression, 138 
Low pressures, compressors, 38 
Lubricating compressors, 182 


Marks and Davis condensed steam 
tables, 193-198 
Measurement of compressed air, 160- 
171 
of draft, 43 
of large quantities of air, 166 
Measuring vacuums, 27 
Mechanical efficiency, 81 
valve of Corliss type, 104 
valves, 98 
Mechanically operated discharge 
valve, I0o 
Mechanics of the fan, 52 
Mercurial air pump, 26 
Meter comparisons, 170 
test results, 171 
Methods of reducing clearance, 71 
Mines and Minerals, 144 
Mixing blower, 127 
Mode of conducting tests, 147 
Modern form of Pitot tube, 162 
Moisture precipitated from air, 3 
Mt. Cenis tunnel, VIIT 
Multi-stage compression, 97 
advantages, 90 


Naperian logarithms, 188-190 


INDEX 


Net efficiency, 81 
Nordberg compressor test, 144 
governor, 10g—110 
intercooler, 92 
Mfg. Co., 109 
Norwalk compressor, 84 
regulator, 108 
Notation of symbols for fan formule, 


47 
Numerical value of R, 10 


Orifice, flow of gas in, 45, 46 
Oxygen in air, I 
in hydraulic compressed air, 137 


Parsons, H. deB., 41-68 
blower, 114 
blades, 115 
Partial pressures, 198 
Peele, Robert, 140 
Perfect gas, characteristic equation, 
IO 
intercooling, 93 
Peripheral speed of fans, 62, 67 
Phenomena of hydraulic air compres- 
sion, 137 
Pipe couplers, 172, 173 
formule, constants, 174, 175 
lines, 171-176 
line formule, 173 
losses, ducts, 48 
Piston, balancing, 123 
-balanced turbo-compressor, 122 
compression, hydraulic, 72 
three-stage diagram, 116 
compressor action, 70 
cross-section, 69 
details, 98-112 
compressors, 69-77 
controlled by multiplicator, 126 
-inlet valve, 102, 108 
Pitot tube, 161, 162 
Pounds of water precipitated per cu. 
ft. cooled air, 3 
Power, 5 
available, 179 
consumed by rotary and piston 
compressors, 52 
for rotary blowers, 51 


205 


Pressures, blower, 50 
oz. per sq. in. in water head to 
inches, 44 
used for various stages, 90 
water column in inches to oz. 
per sq. in., 44 
Proper receiver pressure for multi- 
stage compression, 96 
Propeller fan, 58 
Proportions of fans and housing, 41, 61 
of rotary blowers, 50 
Psychrometers, 192 
Pump, dry air condenser, 26 
Pumps, condenser, 26-30 


R, numerical value, 10 
Radial wheel fan, 58 
Railway and Engineering Review, 130 
Rand Imperial unloader, 111 
Rateau blower, 114 
multiplicator, 125 
turbo-compressor, 128 
Ratio of air cylinder to low-pressure 
steam cylinder, 29 
of air cylinder to volume of con- 
densed steam, 29 
of port to cylinder area, 100 
Real specific heat, 8 
Receiver aftercoolers, 159 
intercoolers, 92 
capacity, 160 
Receivers, 159 
Regulator, belt, 105, 107 
and governor combined, 109 
Norwalk, 108 


Regulators and unloading devices, 
105 

Relation between altitude and volume, 
141 


specific heats, 10 

Relations between P, v and T for 
adiabatic and _ exponentia 
changes, 16 

Relative humidity, 192 

Restricted discharge, 42 

Results of meter tests, 171 

of tests, 148 
Richards, Frank, 175 
Right-angle bend resistance, 49 


206 


Robinson, S. W., 163 

Rotary blowing machines, 49 
blowers, proportions, 50 

Rooms, air required, 40 

Rowse, S. W., 200 

Runners, 119 


Salt evaporating effects, 32 
Sangster, Wm., 39 
Schmidt, Henry F., 81 
Sectional view of Thomas 
meter, 168 
Selection of air compressors, 179-182 
Semi-mechanical valves, 103 
Shape of fan blades, 58, 61, 66, 67 
Simple form of Pitot tube, 161 
Single-stage compression, horse-power 
required, 74 
Sirocco double inlet fan, 68 
Size and type of compressor, 181 
of water and air pumps, 28 
Sketch of meters placed tandem for 
testing, 170 
Sommeiller’s compressor, IX 
Southwork blowing engine valve, 104 
Specific heat, 7 
apparent, and real, 8 
at constant pressure, 7 
at constant volume, 7 
at various pressures 
peratures, 8 
volume of water vapor, 193-1098 
Speed of fans, 58, 62, 67 
of turbo-compressors, 113 
Sperr, aa Ws E36 
Sprengle air pump, 26 
St. John’s meter, 166 
Standards of measurement, 160 
Steam cylinder size, 30 
Steel plate fans, 61, 64 
Straight line compressor, 84 
Stuffing boxes, 123 
Suction line, 73 
Surface of intercoolers, 93 
Sullivan air reheater, 177 
Mch) Cota 153 
Summary of tests, 157 


Syphon, 37 
bulk head, 131 


electric 


and tem- 


INDEX 


Taylor, Charles H., 133 
compressor, 133 
efficiency, 134 
at Ainsworth, B. C., 135 
at Magog, Quebec, 134 
at Victoria mine Michigan, 136 
Temperature, 5 
absolute, 5 
Temperatures due to adiabatic com- 
pression, 22, 23 
Test curves, Jaeger’s turbo-blower, 124 
of hydraulic compressor, 136 
of plant No. 1, 148-151 
of plant No. 2, 151-154 
of plant No. 3, 154-156 
of plant No. 4, 156-157 
Tests, mode of conducting, 147 
Thomas, Cy Ci rr60 
meter, 168 
diagram, 169 
Three-quarter housed steel plate fan, 
64 
Tightness between stages, 120 
Towl, Forrest, M., 160, 171 
Trompe, 129 
True volumetric efficiency, 80 
Turbine blast or Sirocco fan, 67 
Turbo-blower coupling arrangements, 
125 
ol2s 000 Cutt. 121 
of 140,000 cu. ft. capacity, 128 
Turbo-compressor cooling, 115 
design, 113 
diagram, 116 
for mixing air and gas, 128 
Jaeger’s, 119 
Parson’s, 114 
Turbo-compressors, 113-128 
Two-stage compressor cards, 147 
Types of blading, 68 


U. S. Navy pump, 30 
Uncovering port to release clearance 
pressure, 71 
Unloader, clearance, 110, 112 
Rand Imperial, 111 
Unloading devices, 110 
Uses of air at low pressures, 38 
Usual velocity in ducts, 47 


Vacuum cleaners, 36 


concentration of liquors, 34, 35 


manufacture of salt, 32 
measurement, 27 
Valve area, 100 
of discharge, 101 
gear, 179 
in cylinder head, 102 
mechanical, 98 
poppet, Io1 
piston-inlet, 102 
setting, 101 
semi-mechanical, 103 
Valves, area of inlet, 100 
automatic, 100 
classification, 98 
cleaning, 182 
Vapor in air, 1 
Velocity, coefficient of, 43 
of air through ports, ror 
meters, 161 
Ventilation, air required, 39, 40 
Venturi meter, 167 
vacuum pump, 26 
Volumetric efficiency, 77 
apparent, 77 
true, 80 
meters, 160 


Water, air in, 29 


INDEX 207 


Water-cooled turbo-compressor, 117, 
118 
Water measurements, hydraulic com- 
pressor tests, 137 
percipitated from compressed air, 
3 
present in saturated air, 2, 55 
required for intercooler, 94 
Webb, Richard, L., 146 
Weight compensation, 83 
of air, 10 
Westinghouse air pump, 85 
governor, 106 
Wet air pump, 27 
displacement meter, 160 
and dry compression, 74 
Weymouth, Thos. R., 164 
Wheeler combined pump, 27 
condenser pump, 28 
Work, 5 
done by a compressor, 23 
of adiabatic change, 15 
of exponential change, 14 
of isothermal change, 12 
required to move a volume of gas 
56 


Zero, absolute, 6 
Zur Nedden, Franz, 81, 113 


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